Ejector

ABSTRACT

An ejector includes a nozzle for decompressing a fluid in any one state of a gas-liquid state, a liquid state and a super-critical state, and a body portion having a fluid suction port and a mixing and pressurizing portion. The ejector is provided with a suction passage through which a fluid drawn from the fluid suction port flows into the mixing and pressurizing portion. The suction passage is changed such that the fluid drawn from the fluid suction port is decompressed in the suction passage in iso-entropy. Alternatively, the suction passage is changed such that a flow velocity of the fluid flowing into the mixing and pressurizing portion from the suction passage is substantially equal to a flow velocity of the fluid flowing from a jet port of the nozzle into the mixing and pressurizing portion, or is equal to or larger than the sound velocity.

CROSS REFERENCE TO RELATED APPLICATION

This application is based on Japanese Patent Applications No.2008-062142 filed on Mar. 12, 2008, No. 2008-062143 filed on Mar. 12,2008, No. 2008-135077 filed on May 23, 2008, No. 2008-135076 filed onMay 23, 2008, and No. 2009-010645 filed on Jan. 21, 2009, the contentsof which are incorporated herein by reference in its entirety.

FIELD OF THE INVENTION

The present invention relates to an ejector configured to draw a fluidby a jet flow of a high-speed fluid jetted from a nozzle. For example,the ejector can be suitably used for a refrigeration cycle device.

BACKGROUND OF THE INVENTION

Conventionally, an ejector is known, which includes a nozzle fordecompressing and expanding a high-pressure fluid, and is configured todraw a fluid from a fluid suction port by a suction action of a jet flowof a high-speed fluid jetted from the nozzle. In the ejector, the jetfluid from the nozzle and the suction fluid from the fluid suction portare mixed in a mixing portion, and the pressure of the mixed fluid isincreased in a diffuser portion by converting the kinetic energy of themixed fluid to the pressure energy of the mixed fluid. Therefore, thepressure of the fluid flowing out of the outlet of the ejector isincreased more than the pressure of the suction fluid.

In an ejector described in JP 2004-340136A (corresponding to US2004/0206111 A1), a passage sectional area at an inlet side of a suctionpassage through which a suction fluid introduced from a fluid suctionport flows into a mixing portion of the ejector is set equal to orlarger than a passage sectional area of the fluid suction port.Therefore, the pressure loss, caused when the suction fluid is drawnfrom the fluid suction port, can be reduced, and the flow amount of thesuction fluid flowing from the fluid suction port can be increased,thereby improving ejector efficiency ηe that is an energy convertingefficiency in the ejector.

In an ejector for a refrigeration cycle device described in JP2003-14318A (corresponding to US 2002/0000095A1), an expanding angle ofa passage wall surface of a diffuser portion is suitably set in an axialsection including the center axial of a nozzle so that a pressurizingamount in the diffuser portion is increased, thereby improving theejector efficiency ηe.

In another ejector described in JP 2004-116807A, a passage wall surfaceof a diffuser portion is formed into a smoothly covered line in an axialsection including the center axial of a nozzle so that an energy losssuch as a scroll flow loss in the diffuser portion can be restricted,thereby improving the ejector efficiency ηe.

The ejector efficiency ηe is defined as in the following formula (F1).

e=(1+Ge/Gnoz)×(ΔP/ρ)/Δi   (F1)

Here, Ge is the flow amount of the suction fluid, Gnoz is the flowamount of the jet fluid, ΔP is the pressurizing amount in the diffuserportion, ρ is the density of the suction fluid, and Δi is the enthalpydifference between the inlet and the outlet of the nozzle.

However, JP 2004-340136A does not describe regarding the pressure losson a downstream side in the suction passage downstream of the fluidsuction port. If the pressure loss in the suction passage changes, theflow amount of the suction fluid or the flow velocity of the fluidflowing into the mixing portion through the suction passage is changed.In addition, when the fluid flowing through the mixing portion and thediffuser portion is in a gas-liquid two-phase state, the inertial forcebecomes different in the gas fluid and the liquid fluid due to thedensity difference between the gas fluid and the liquid fluid, andthereby it is difficult to uniformly mix the jet fluid and the suctionfluid in the mixing portion of the ejector.

Thus, in the diffuser portion of the ejector, the kinetic energy of thefluid is converted to the pressure energy in an inhomogeneous state, andthereby the ejector efficiency ηe cannot be sufficiently improved. Here,the inhomogeneous state means a state other than a homogeneous statethat includes a complete gas state, a complete liquid state and ahomogeneously mixed state in which the gas fluid and the liquid fluidare homogeneously mixed with approximately the same flow velocity. In anexample of the inhomogeneously mixed sate of the gas fluid and theliquid fluid, the flow velocity of the gas fluid is different from theflow velocity of the liquid fluid.

Furthermore, in JP 2003-14318A or JP 2004-116807A, the ejector isconfigured to improve the ejector efficiency ηe, in a case where thefluid of the homogeneous state passes through the mixing portion and thediffuser portion of the ejector. Actually, it is difficult for thegas-liquid two-phase fluid passing through the mixing portion and thediffuser portion of the ejector to be in the homogeneous state.Accordingly, when gas-liquid two-phase refrigerant passes through themixing portion and the diffuser portion in the ejector, it is difficultto sufficiently improve the ejector efficiency ηe.

SUMMARY OF THE INVENTION

In view of the foregoing problems, it is an object of the presentinvention to sufficiently improve the ejector efficiency ηe in anejector having a mixing and pressurizing portion in which the kineticenergy of a gas-liquid two-phase fluid is converted to the pressureenergy thereof.

It is another object of the present invention to provide an ejectorprovided with a suction passage which is configured to improve theejector efficiency ηe.

The following aspects of the present invention are devised by theinventors of the present application based on the following experimentsand studies. An ejector recovers the energy lost in decompression andexpansion by decompressing and expanding a fluid in iso-entropy at anozzle, and converts the recovered energy (recovery energy) to thepressure energy, so as to improve the ejector efficiency ηe.

If it is possible for all the recovery energy to be converted to thepressure energy, the ejector efficiency ηe will be made maximum. Theinventors of the present application examined and studied in detailregarding the recovery energy used actually in the ejector. That is, theenergy capable of being used for pressurizing the fluid, among therecovery energy, is studied.

FIG. 28 shows results examined and studied by the inventors of thepresent application. In an ejector of a comparative example having amixing portion and a diffuser portion shown in FIG. 29, a total recoveryenergy at an inlet of a mixing portion can be divided into E1 to E4 asshown in FIG. 28. In FIG. 28, E1 indicates the energy used forpressurizing, E2 indicates a remaining kinetic energy without beingused, E3 indicates energy transmission loss, and E4 indicates the otherloss. As shown from FIG. 28, the energy E1 used for pressurizing isabout 20% of the total recovery energy, and the other energy E2, E3, E4is not used for pressurizing. The remaining kinetic energy E2 isremained as a flow velocity of the fluid flowing out of the diffuserportion of the ejector without being converted to the pressure energy.

The energy transmission loss E3 includes the energy transmission losscaused by transmitting the kinetic energy of the liquid fluid to the gasfluid, while the liquid fluid and the gas refrigerant pass through thediffuser portion of the ejector, for example. As shown in FIG. 28, theratio of the energy transmission loss E3, in the energy E2, E3 and E4without being used for pressurizing, is relatively large, as comparedwith the energy E1 used for pressurizing.

The inventors of the present application studied regarding the reductionof the energy transmission loss E3 between the gas fluid and the liquidfluid. When the energy transmission loss E3 between the gas fluid andthe liquid fluid is reduced and is used for the pressurizing, theejector efficiency ηe can be effectively improved. Thus, the inventorsperformed experiments for effectively transmitting energy from theliquid fluid having a high flow velocity than that of the gas fluid, tothe gas fluid.

In a case of a free fall rigid body, the flow velocity in a verticaldownward direction is increased by acceleration of gravity. Then, theflow velocity of the free fall rigid body is reached to a certainterminal velocity in accordance with a balance with resistance receivedfrom the circumference air.

That is, the flow velocity of the free fall rigid body is not increasedmore than the terminal velocity after reaching the terminal velocity.Therefore, the flow velocity of the free fall rigid body becomes maximumwhen reaching to the terminal velocity. It means that the kinetic energyof the rigid body can be rapidly transmitted to the circumference airwhen the rigid body rapidly reaches to the terminal velocity. In FIG.29, the liquid fluid grain (i.e., virtual liquid particle) passingthrough the diffuser portion is supposed as the rigid body, and the gasfluid passing through the diffuser portion is supposed as thecircumference air. In the supposed state of FIG. 29, the inventors ofthe present application studied regarding an effective energytransmission between the liquid fluid and the gas fluid passing throughthe diffuser portion.

The upper part of FIG. 29 is a graph showing variation in velocity ofthe gas fluid and velocity of the liquid fluid within an ejector. Thesolid line LA1 shows a variation in the liquid fluid (e.g., liquidrefrigerant) in the ejector of the comparison example, and the solidline GA1 shows a variation in the gas fluid (e.g., gas refrigerant) inthe ejector of the comparison example. As shown in the solid lines LA1and GA1 of FIG. 29, the flow velocity of the gas fluid is greatly fasterthan that of the liquid fluid in a nozzle of the ejector in thecomparison example by the difference in the inertial force due to thedensity difference between the gas fluid and the liquid fluid. Thus, inthe mixed fluid of the jet fluid and the suction fluid flowing into amixing portion of the ejector, the flow velocity of the gas fluidbecomes faster than the flow velocity of the liquid refrigerant.

The grains of the liquid fluid flowing into the mixing portion areaccelerated together with the circumference gas fluid, and then the flowvelocity of the grains of the liquid fluid becomes equal to the flowvelocity of the gas fluid. After the flow velocity of the grains of theliquid fluid becomes equal to the flow velocity of the gas fluid, theflow velocity of the liquid fluid is not more accelerated, and reachesto the terminal velocity.

The flow velocity of the grain of the liquid fluid after reaching to theterminal velocity is reduced while applying a force corresponding to theresistance force to the circumferential gas fluid as a reaction force.At this time, the kinetic amount is transmitted from the grains of theliquid fluid to the gas fluid, and the total value of impulses appliedfrom the grains of the liquid fluid to the gas fluid becomes thepressurizing amount (pressure energy) of the gas fluid.

Accordingly, if the grains of the liquid fluid flowing into the mixingportion of the ejector are rapidly reached to the terminal velocity, thekinetic energy included in the liquid fluid can be rapidly transmittedto the gas fluid. Thus, after the flow velocity of the liquid fluidreaches to the terminal velocity, the kinetic energy of the liquid fluidcan be effectively transmitted to the gas fluid. Furthermore, when theterminal velocity itself of the grains of the liquid fluid is increased,the pressurizing amount of the gas fluid can be increased, therebyimproving the ejector efficiency ηe.

In FIG. 29, the chain line LA2 indicates a variation in the flowvelocity of the liquid fluid of an ejector according to an example ofthe present invention, and the chain line GA2 indicates a variation inthe flow velocity of the gas fluid of the ejector according to theexample of the present invention. As shown by the chain lines LA2 andGA2 in FIG. 29, when the flow velocity of the gas fluid flowing into themixing portion is increased, the terminal velocity of the grains of theliquid fluid can be increased, as compared with the comparison exampleshown by the solid lines LA1 and GA1. Thus, in the example of thepresent invention shown by the chain lines LA2 and GA2 in FIG. 29,because a large amount of the kinetic energy can be converted to thepressure energy, the energy transmission loss between the gas fluid andthe liquid fluid can be effectively reduced, thereby significantlyimproving the ejector efficiency ηe.

According to an aspect of the present invention, an ejector includes anozzle configured to decompress and expand a fluid in any one state of agas-liquid two-phase state, a liquid state and a super-critical state,and a body portion in which the nozzle is disposed. The body portion hasa fluid suction port from which a fluid is drawn by a jet flow of thefluid jetted from a jet port of the nozzle, and a mixing andpressurizing portion in which the fluid jetted from the jet port of thenozzle and the fluid drawn from the fluid suction port are mixed andkinetic energy of the mixed fluid in a gas-liquid two-phase state isconverted to pressure energy thereof. The ejector is provided with asuction passage through which the fluid drawn from the fluid suctionport flows into an inlet of the mixing and pressurizing portion, and afluid passage area of the suction passage is configured to be changedsuch that the fluid drawn from the fluid suction port is decompressed inthe suction passage substantially in iso-entropy.

Accordingly, the energy loss while the suction fluid passes through thesuction passage can be reduced. Thus, the flow velocity of the fluidflowing into the mixing and pressurizing portion from the suctionpassage can be increased, thereby increasing the flow velocity of thegas fluid flowing into the mixing and pressurizing portion. As a result,the terminal velocity of grains of the liquid fluid flowing into themixing and pressurizing portion can be increased, and the pressurizingamount in the gas fluid can be increased in the ejector even when thekinetic energy of the gas-liquid two-phase fluid is converted to thepressure energy thereof in the mixing and pressurizing portion.Therefore, the ejector efficiency can be effectively improved.

According to another aspect of the present invention, an ejectorincludes a nozzle configured to decompress and expand a fluid in any onestate of a gas-liquid two-phase state, a liquid state and asuper-critical state, and a body portion in which the nozzle isdisposed. The body portion has a fluid suction port from which a fluidis drawn by a jet flow of the fluid jetted from a jet port of thenozzle, and a mixing and pressurizing portion in which the fluid jettedfrom the jet port of the nozzle and the fluid drawn from the fluidsuction port are mixed and kinetic energy of the mixed fluid in agas-liquid two-phase state is converted to pressure energy thereof. Theejector is provided with a suction passage through which the fluid drawnfrom the fluid suction port flows into an inlet of the mixing andpressurizing portion. In the ejector, a fluid passage area of thesuction passage is configured to be changed such that a flow velocity ofthe fluid flowing into the mixing and pressurizing portion from thesuction passage is substantially equal to a flow velocity of the fluidflowing from the jet port of the nozzle into the mixing and pressurizingportion. As a result, the terminal velocity of grains of the liquidfluid flowing into the mixing and pressurizing portion can be increased,and the pressurizing amount in the gas fluid can be increased in theejector even when the kinetic energy of the gas-liquid two-phase fluidis converted to the pressure energy thereof in the mixing andpressurizing portion. Therefore, the ejector efficiency can beeffectively improved. Here, the meaning of “substantially equal”includes that the flow velocity of the fluid flowing into the mixing andpressurizing portion from the suction passage completely corresponds toor slightly different from the flow velocity of the fluid flowing fromthe jet port of the nozzle into the mixing and pressurizing portion.

According to another aspect of the present invention, an ejectorincludes a nozzle configured to decompress and expand a fluid in any onestate of a gas-liquid two-phase state, a liquid state and asuper-critical state, and a body portion in which the nozzle isdisposed. The body portion has a fluid suction port from which a fluidis drawn by a jet flow of the fluid jetted from a jet port of thenozzle, and a mixing and pressurizing portion in which the fluid jettedfrom the jet port of the nozzle and the fluid drawn from the fluidsuction port are mixed and kinetic energy of the mixed fluid in agas-liquid two-phase state is converted to pressure energy thereof. Theejector is provided with a suction passage through which the fluid drawnfrom the fluid suction port flows into an inlet of the mixing andpressurizing portion. Furthermore, a fluid passage area of the suctionpassage is configured to be changed such that a flow velocity of thefluid flowing into the mixing and pressurizing portion from the suctionpassage is equal to or larger than a sound velocity. As a result, theterminal velocity of grains of the liquid fluid flowing into the mixingand pressurizing portion can be increased, and the pressurizing amountin the gas fluid can be increased in the ejector even when the kineticenergy of the gas-liquid two-phase fluid is converted to the pressureenergy thereof in the mixing and pressurizing portion.

In any one aspect of the present invention, the fluid passage area ofthe suction passage may be gradually reduced toward downstream in a flowdirection of the fluid flowing in the suction passage. In this case, areduce degree of the fluid passage area at an inlet side of the suctionpassage may be larger than a reduce degree of the fluid passage area atan outlet side of the suction passage.

Alternatively, the fluid passage area of the suction passage at an inletside of the suction passage may be gradually reduced toward downstreamin the flow direction of the fluid flowing in the suction passage, andthe fluid passage area of the suction passage at an outlet side of thesuction passage may be gradually increased toward downstream in the flowdirection of the fluid flowing in the suction passage.

The suction passage may be provided between an outer peripheral surfaceof the nozzle and an inner peripheral surface of the body portion, ormay be configured by another nozzle to be provided therein.Alternatively, the nozzle and the suction passage may be configured,such that an enthalpy difference (ΔH) between enthalpy of the fluid atan inlet of the nozzle and enthalpy of the fluid at the jet port of thenozzle is equal to or larger than an enthalpy difference (Δh) betweenenthalpy of the fluid at the inlet of the suction passage and enthalpyof the fluid at the outlet of the suction passage.

According to another aspect of the present invention, an ejectorincludes a nozzle configured to decompress and expand a fluid in any onestate of a gas-liquid two-phase state, a liquid state and asuper-critical state, and a body portion in which the nozzle isdisposed. The body portion has a fluid suction port from which a fluidis drawn by a jet flow of the fluid jetted from a jet port of thenozzle, and a mixing and pressurizing portion in which the fluid jettedfrom the jet port of the nozzle and the fluid drawn from the fluidsuction port are mixed and kinetic energy of the mixed fluid in agas-liquid two-phase state is converted to pressure energy thereof. Themixing and pressurizing portion is configured by a straight portionextending from the inlet of the mixing and pressurizing portion in arange, and an expanding portion extending from a downstream end of thestraight portion to the outlet of the mixing and pressurizing portion.The straight portion is cylindrical passage having a constant passagearea in its entire range, and the expending portion is configured suchthat a passage sectional area of the expanding portion is graduallyincreased toward downstream in a flow direction of the fluid. As aresult, the terminal velocity of grains of the liquid fluid flowing intothe mixing and pressurizing portion can be increased, and thepressurizing amount in the gas fluid can be increased in the ejectoreven when the kinetic energy of the gas-liquid two-phase fluid isconverted to the pressure energy thereof in the mixing and pressurizingportion.

For example, the range of the straight portion may be set such that theflow velocities of gas fluid and liquid fluid within the fluid flowinginto the mixing and pressurizing portion become equal to each other inthe range. Alternatively, when a length of the straight portion in anaxial direction of the nozzle is L1 and a length from the inlet of themixing and pressurizing portion to the outlet of the mixing andpressurizing portion in the axial direction is L2, the mixing andpressurizing portion is configured such that 0<L1/L2≦0.4. Furthermore,the mixing and pressurizing portion may be configured such that thefluid is pressurized in iso-entropy in the mixing and pressurizingportion.

In the ejector, a sectional shape of a wall surface of the expandingportion in a section including an axial line of the nozzle may be astraight line or a curved line. Alternatively, the sectional shape ofthe wall surface of the expanding portion in a section including theaxial line of the nozzle may be formed by combining plural straightlines or may be formed by combining at least a straight line and acurved line. Alternatively, an expanding degree of the expanding portionat an inlet side of the expanding portion may be larger than anexpanding degree of the expanding portion at an outlet side of theexpanding portion.

BRIEF DESCRIPTION OF THE DRAWINGS

Additional objects and advantages of the present invention will be morereadily apparent from the following detailed description of preferredembodiments when taken together with the accompanying drawings. Inwhich:

FIG. 1 is a schematic diagram showing a refrigeration cycle devicehaving an ejector according to a first embodiment of the presentinvention;

FIG. 2A is an axial sectional view of the ejector including an axialline of a nozzle according to the first embodiment, FIG. 2B is across-sectional view taken along the line IIB-IIB of FIG. 2A, and FIG.2C is a cross-sectional view taken along the line IIC-IIC of FIG. 2A;

FIG. 3 is a graph showing a variation in a ratio of a refrigerantpassage sectional area of a suction passage to a refrigerant passagesectional area at an inlet of the suction passage, in the ejectoraccording to the first embodiment;

FIG. 4 is a schematic diagram showing a passage configuration of amixing and pressurizing portion of the ejector according to the firstembodiment;

FIG. 5A is a Mollier diagram showing a refrigerant state in arefrigerant cycle of the refrigeration cycle device according to thefirst embodiment, and FIG. 5B is an enlarged view showing the part VB inFIG. 5A;

FIG. 6 is a graph showing variations in the flow velocity of gasrefrigerant and the flow velocity of liquid refrigerant in the ejectorof the first embodiment and in an ejector of a comparison example;

FIG. 7A is a graph showing variations in a flow velocity of refrigerantand a pressurizing amount (ΔP) in the ejector according to the firstembodiment, and FIG. 7B is a graph showing variations in a flow velocityof refrigerant and a pressurizing amount (ΔP) in the ejector accordingto a comparison example;

FIG. 8 is a graph showing an energy amount (E1) to be used forpressurizing, a remain kinetic energy (E2), an energy transmission loss(E3) and the other loss (E4), according to the first embodiment and thecomparison example;

FIG. 9 is a graph showing a variation in a ratio of a refrigerantpassage sectional area of a suction passage to a refrigerant passagesectional area at an inlet of the suction passage, in an ejectoraccording to a second embodiment of the present invention;

FIG. 10 is a graph showing a variation in a ratio of a refrigerantpassage sectional area of a suction passage to a refrigerant passagesectional area at an inlet of the suction passage, in an ejectoraccording to a third embodiment of the present invention;

FIG. 11 is a schematic diagram showing a passage configuration of amixing and pressurizing portion in an ejector according to a fourthembodiment of the present invention;

FIG. 12 is a schematic diagram showing a passage configuration of amixing and pressurizing portion in an ejector according to a fifthembodiment of the present invention;

FIG. 13 is a schematic diagram showing a passage configuration of amixing and pressurizing portion in an ejector according to a sixthembodiment of the present invention;

FIG. 14 is a schematic diagram showing a passage configuration of amixing and pressurizing portion in an ejector according to a seventhembodiment of the present invention;

FIG. 15 is a schematic diagram showing a passage configuration of amixing and pressurizing portion in an ejector according to an eighthembodiment of the present invention;

FIG. 16 is an axial sectional view showing an ejector according to aninth embodiment of the present invention;

FIG. 17 is an axial sectional view showing an ejector according to atenth embodiment of the present invention;

FIG. 18 is a Mollier diagram showing a refrigerant state in arefrigerant cycle of a refrigeration cycle device according to aneleventh embodiment of the present invention;

FIG. 19 is a Mollier diagram showing a refrigerant state in arefrigerant cycle of a refrigeration cycle device according to a twelfthembodiment of the present invention;

FIG. 20 is a Mollier diagram showing a refrigerant state in arefrigerant cycle of a refrigeration cycle device according to athirteenth embodiment of the present invention;

FIG. 21 is a Mollier diagram showing another refrigerant state in therefrigerant cycle of the refrigeration cycle device according to thethirteenth embodiment of the present invention;

FIG. 22 is a schematic diagram showing a refrigeration cycle devicehaving an ejector according to a fourteenth embodiment of the presentinvention;

FIG. 23 is a Mollier diagram showing a refrigerant state in arefrigerant cycle of the refrigeration cycle device according to thefourteenth embodiment of the present invention;

FIG. 24 is a schematic diagram showing a refrigeration cycle devicehaving an ejector according to a fifteenth embodiment of the presentinvention;

FIG. 25 is a Mollier diagram showing a refrigerant state in arefrigerant cycle of the refrigeration cycle device according to thefifteenth embodiment of the present invention;

FIG. 26 is a schematic diagram showing a refrigeration cycle devicehaving an ejector according to another embodiment of the presentinvention;

FIG. 27A is a Mollier diagram showing a refrigerant state in arefrigerant cycle of a refrigeration cycle device according to anotherembodiment of the present invention, and FIG. 27B is a Mollier diagramshowing a refrigerant state in a refrigerant cycle of a refrigerationcycle device according to another embodiment of the present invention;

FIG. 28 is a graph showing energy division in a recovery energy at aninlet of a mixing portion of an ejector in a comparison example; and

FIG. 29 is a graph showing experimental results in velocity distributionof gas fluid and liquid fluid in an ejector.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS First Embodiment

A first embodiment of the present invention will be described withreference to FIGS. 1 to 8. In the first embodiment, an ejector 16 of thepresent invention is typically used for a refrigeration cycle device 10shown in FIG. 1. The refrigeration cycle apparatus 10 shown in FIG. 1can be used for a vehicle air conditioner, for example.

In the refrigeration cycle device 10, a compressor 11 is configured todraw refrigerant, to compress the drawn refrigerant, and to dischargethe compressed high-pressure and high-temperature refrigerant. Thecompressor 11 is driven and rotated by a vehicle engine (not shown) viaan electromagnetic clutch and a belt, or the like, as an example.

The compressor 11 may be a variable displacement compressor in which adischarge capacity of the refrigerant can be continuously adjustable, ormay be a fixed displacement compressor in which the discharge capacityof the refrigerant can be adjusted by changing a compressor operationratio. For example, in the fixed displacement compressor, the compressoroperation ratio is changed by interruption of the electromagneticclutch. Alternatively, an electrical compressor may be used as thecompressor 11 such that the refrigerant discharge capacity of thecompressor 11 can be adjusted by adjusting a rotation speed of anelectrical motor.

A refrigerant radiator 12 used as a heat exchanger for heat radiationsuch as a refrigerant cooler is disposed at a refrigerant discharge sideof the compressor 11. The radiator 12 is configured to perform a heatexchange between the high-pressure refrigerant discharged from thecompressor 11 and outside air (i.e., air outside a vehicle compartment)blown by a blower fan 12 a, thereby cooling the high-pressurerefrigerant in the radiator 12.

As the refrigerant used in a refrigerant cycle of the refrigerationcycle apparatus 10, a Freon-based refrigerant such as HFC134a may beused so that a refrigerant pressure on a high-pressure side in therefrigerant cycle does not excess the critical pressure of therefrigerant. In this case, the radiator 12 is used as a condenser inwhich the refrigerant is cooled and condensed therein.

A receiver 12 b is located at a refrigerant outlet side of the radiator12. The receiver 12 b is a gas-liquid separator with a verticallyelongated tank. The receiver 12 b is configured to separate therefrigerant flowing therein into gas refrigerant and liquid refrigerant,and to store therein surplus liquid refrigerant in the refrigerantcycle. The receiver 12 b has a liquid refrigerant outlet at a lower sideof the tank so that the liquid refrigerant flows out of the receiver 12from the liquid refrigerant outlet. As an example of the presentembodiment, the receiver 12 b is formed integrally with the radiator 12,as shown in FIG. 1. However, the receiver 12 b may be located separatelyfrom the radiator 12, or may be omitted.

A branch portion 13 is connected to the liquid refrigerant outlet of thereceiver 12 b, and is configured to divide the liquid refrigerant fromthe receiver 12 b into two streams. For example, the branch portion 13is a three-way joint member having one refrigerant inlet and first andsecond refrigerant outlets. The three-way joint member used as thebranch portion 13 may be configured by bonding pipes having differentpipe diameters, or may be configured by providing plural refrigerantpassages in a metal block member or a resin block member.

One of the refrigerant streams branched at the branch portion 13 flowsinto a first refrigerant passage 14 a (nozzle-side refrigerant passage),and the other one of the refrigerant streams branched at the branchportion 13 flows into a second refrigerant passage 14 b (suction-siderefrigerant passage). One end of the first refrigerant passage 14 a isconnected to the first refrigerant outlet of the branch portion 13, andthe other end of the first refrigerant passage 14 a is connected to aninlet of a nozzle 16 a of the ejector 16, so that one of the refrigerantstreams branched at the branch portion 13 flows into the nozzle 16 athrough the first refrigerant passage 14 a. One end of the secondrefrigerant passage 14 b is connected to the second refrigerant outletof the branch portion 13, and the other end of the second refrigerantpassage 14 b is connected to a refrigerant suction port 16 d of theejector 16, so that the other one of the refrigerant streams branched atthe branch portion 13 flows into the refrigerant suction port 16 dthrough the second refrigerant passage 14 b.

An expansion valve 15 is located in the first refrigerant passage 14 aat an upstream side of the nozzle 16 a of the ejector 16 in arefrigerant flow of the first refrigerant passage 14 a. The expansionvalve 15 is used as a decompression portion configured to decompresshigh-pressure liquid refrigerant flowing from the receiver 12 b into thefirst refrigerant passage 14 a to be in a gas-liquid two-phase statehaving a middle pressure. The expansion valve 15 is also used as a flowamount adjusting portion for adjusting a flow amount of the refrigerantflowing into the nozzle 16 a.

In the present embodiment, the expansion valve 15 is a thermal expansionvalve having a temperature sensing portion 15 a that is located at arefrigerant suction passage of the compressor 11 so as to detect asuper-heat degree of the refrigerant to be drawn into a refrigerantsuction side of the compressor 11. In this embodiment, the refrigeranton the refrigerant suction side of the compressor 11 corresponds to therefrigerant at a refrigerant outlet side of a first evaporator 17. Thatis, the temperature sensing portion 15 a detects the super-heat degreeof the refrigerant at a refrigerant outlet side of the first evaporator17 based on at least one of temperature and pressure of the refrigerantat the refrigerant outlet side of the first evaporator 17, and a valveopen degree of the expansion valve 15 is adjusted using a mechanicalmechanism or an electrical mechanism so that the super-heat degree ofthe refrigerant at the refrigerant outlet of the first evaporator 17 isapproached to a predetermined value. Thus, the flow amount of therefrigerant flowing to downstream of the expansion valve 15 can beadjusted.

The other throttle structure or decompression device may be used insteadof the thermal expansion valve 15. For example, a decompression devicesuch as an electrical variable throttle device or a fixed throttledevice, or the other type expansion valve may be used instead of thethermal expansion device 15.

The elector 16 is located at a refrigerant outlet side of the expansionvalve 15. The ejector 14 is adapted as a decompression portion forfurther decompressing the refrigerant flowing from the expansion valve15, and as a refrigerant circulation portion for circulating therefrigerant by the suction action of a high-speed refrigerant flowjetted from the nozzle 16 a. The structure of the ejector 16 will be nowdescribed with reference to FIGS. 2A to 4.

FIG. 2A is an axial sectional view of the ejector 16 taken along asection including an axial line, FIG. 2B is a cross-sectional view takenalong the line IIB-IIB in FIG. 2A at an inlet of a suction passage 16iof the ejector 16, and FIG. 2C is a cross-sectional view taken along theline IIC-IIC in FIG. 2A at an outlet of the suction passage 16i of theejector 16.

As shown in FIG. 2A, the ejector 16 is configured by the nozzle 16 a anda body portion 16 b. The nozzle 16 a is made of a metal such as astainless alloy, and is formed into an approximately cylindrical shapehaving a taper end portion tapered toward downstream in the refrigerantflow. The refrigerant passage sectional area of the nozzle 16 a ischanged in the refrigerant flow direction so that the refrigerantflowing into the nozzle 16 a is decompressed and expanded iniso-entropy.

A refrigerant jet port 16 c, from which the refrigerant is jetted fromthe nozzle 16 a, is formed at a tip end of the taper end portion of thenozzle 16 a. The nozzle 16 a is disposed in the body portion 16 b and isattached into the body portion 16 b such that the refrigerant isprevented from being leaked from a fixing portion of the nozzle 16 a andthe body portion 16 b. For example, the nozzle 16 a may be air-tightlyfitted into the body portion 16 b, or may be air-tightly bonded to thebody portion 16 b by using a bonding means such as welding, pressing orbrazing, or the like.

For example, the nozzle 16 a may be a Laval nozzle having a throatportion at which the passage sectional area becomes smallest within therefrigerant passage inside of the nozzle 16 a. The nozzle 16 a isconfigured such that the flow velocity of the refrigerant jetted fromthe refrigerant jet port 16 c of the nozzle 16 a becomes equal to orlarger than the sound velocity. Alternatively, the nozzle 16 a may beconfigured by a taper nozzle so that the flow velocity of therefrigerant jetted from the refrigerant jet port 16 c of the nozzle 16 abecomes equal to or larger than the sound velocity.

The body portion 16 b can be made of a metal, for example, aluminum oran aluminum alloy, or may be made of a material other than the metalsuch as resin. The body portion 16 b is provided with the refrigerantsuction port 16 b penetrating through the interior and the exterior ofthe body portion 16 b in a radial direction perpendicular to the axialdirection of the nozzle 16 a. The refrigerant suction port 16 b is openin the body portion 16 b at a portion radially outside of the nozzle 16a. The body portion 16 b has therein a mixing and pressurizing portion16 e extending in the axial direction (longitudinal direction) from aposition of the refrigerant jet port 16 c to the refrigerant outlet(downstream end).

The refrigerant suction port 16 d is coupled to a refrigerant outletside of a second evaporator 19 so that the refrigerant from the secondevaporator 19 is drawn into the suction passage 16 i from therefrigerant suction port 16 d. The refrigerant suction port 16 d isprovided at an outer peripheral side of the nozzle 16 a, andcommunicates with a space at the refrigerant jet port 16 c within thebody portion 16 b through the suction passage 16 i.

An inlet space, into which refrigerant from the refrigerant suction port16 d flows, is provided within the body portion 16 b around therefrigerant suction port 16 d, and the suction passage 16 i is providedbetween the outer wall surface of the taper end portion of the nozzle 16a and an inner wall surface of the body portion 16 b. Therefore, therefrigerant flowing from the refrigerant suction port 16 d into theinlet space of the body portion 16 b is introduced into an inlet of themixing and pressurizing portion 16 e via the suction passage 16 i. Here,the inlet of the mixing and pressurizing portion 16 e in the bodyportion 16 b substantially corresponds to the position of therefrigerant jet port 16 c of the nozzle 16 a in the axial direction.

FIG. 2B shows a refrigerant passage sectional area Ain at the inlet ofthe suction passage 16 i, and FIG. 2C shows a refrigerant passagesectional area Aout at the outlet of the suction passage 16 i. As shownin FIGS. 2B and 2C, refrigerant passage area Aout at the outlet of thesuction passage 16 i is smaller than the refrigerant passage area Ain atthe inlet of the suction passage 16 i.

FIG. 3 shows a variation in a ratio (passage area ratio) of arefrigerant passage sectional area of the suction passage 16 i in therefrigerant flow direction, to the refrigerant passage sectional area atthe inlet of the suction passage 16 i. As shown by the solid line inFIG. 3, the passage area ratio of the suction passage 16 i is graduallyreduced from the inlet to the outlet of the suction passage 16 i in therefrigerant flow of the suction passage 16 i. As shown in FIG. 3, areduce degree of the passage sectional area on a side of the inlet ofthe suction passage 16 i is larger than a reduce degree of the passagesectional area on a side of the outlet of the suction passage 16 i.

Specifically, as shown by the solid-line graph of FIG. 3, the passagesectional area of the suction passage 16 i is rapidly reduced in a rangefrom the inlet of the suction passage 16 i approximately to a middleposition of the suction passage 16 i, and the passage sectional area ofthe suction passage 16 i is slowly reduced approximately from the middleposition of the suction passage 16 i to the outlet of the suctionpassage 16 i. Thus, as compared with the comparative chain linestraightly connecting the inlet and the outlet of the suction passage 16i, the variation line (i.e., solid line in FIG. 3) of the passage arearatio of the suction passage 16 i is positioned under the comparativechain line and is made convex downwardly.

In the present embodiment, the passage sectional area of the suctionpassage 16 i is changed as described above so that the flow velocity ofthe refrigerant passing through the suction passage 16 i becomes equalto or greater than the sound velocity. Thus, the flow velocity of thesuction refrigerant flowing into the mixing and pressurizing portion 16e from the suction passage 16 i can be made approximately equal to theflow velocity of the jet flow jetted from the jet port 16 c of thenozzle 16 a into the mixing and pressurizing portion 16 e. Accordingly,it is possible for the suction refrigerant can be decompressed iniso-entropy in the suction passage 16 i.

As shown in FIG. 2A, the mixing and pressurizing portion 16 e ispositioned just downstream of the nozzle 16 a and the suction passage 16i, so that the kinetic energy of gas-liquid two-phase refrigerant isconverted to the pressure energy thereof in the mixing and pressurizingportion 16 e while the jet refrigerant jetted from the nozzle 16 a andthe suction refrigerant drawn from the refrigerant suction port 16 d aremixed in the mixing and pressurizing portion 16 e.

The mixing and pressurizing portion 16 e is configured by a straightportion 16 g from the inlet of the mixing and pressurizing portion 16 eto a predetermined range, and an expanding portion 16 h from thedownstream side of the straight portion 16 g to the outlet of theejector 16. The straight portion 16 g of the mixing and pressurizingportion 16 e is a cylindrical passage having a constant passagesectional area. The expanding portion 16 h is gradually enlarged in thepassage sectional area from its inlet toward downstream.

The straight portion 16 g is provided in a range from the inlet of themixing and pressurizing portion 16 e, such that the flow velocity of thegas refrigerant and the flow velocity of the liquid refrigerant in therefrigerant flowing into the mixing and pressurizing portion 16 e becomeapproximately equal to each other in the straight portion 16 e. When thelength of the straight portion 16 g in the axial direction of the nozzle16 a is L1 and when the length of the mixing and pressurizing portion 16e in the axial direction of the nozzle 16 a is L2, a ratio L1/L2 is setabout 0.2, as an example.

The refrigerant passage shape of the expanding portion 16 h in thesection including the center line (axial line) is changed in a smoothlycurved line as shown in FIG. 4. An increase degree in the refrigerantpassage sectional area of the expanding portion 16 h is changed as shownin FIG. 4. The increase degree on an inlet side of the refrigerantpassage sectional area of the expanding portion 16 h is larger than theincrease degree on an outlet side of the refrigerant passage sectionalarea of the expanding portion 16 h. That is, the increase degree on theinlet side of the refrigerant passage sectional area of the expandingportion 16 h is relatively rapidly increased, and the increase degree onthe outlet side of the refrigerant passage sectional area of theexpanding portion 16 h is relatively slowly increased, as compared withthe mean increase degree from the inlet to the outlet of the expandingportion 16 h.

As shown in FIG. 4, the passage wall surface of the expanding portion 16h on the section including the axial line of the expanding portion 16 h,the sectional shape at the inlet side of the passage wall surface of theexpanding portion 16 h is formed into a curved line 101 with a slightconvex toward an inner peripheral side, and the sectional shape at theoutlet side of the passage wall surface of the expanding portion 16 h isformed into a curved line 102 with a slight convex toward an outerperipheral side. The straight portion 16 g and the expanding portion 16h of the mixing and pressurizing portion 16 e are continuously extended,and are configured such that the refrigerant is substantiallypressurized in iso-entropy in the mixing and pressurizing portion 16 ewhile the refrigerant is prevented from being separated from the passagewall surface of the mixing and pressurizing portion 16 e at the outletof the mixing and pressurizing portion 16 e.

Thus, the energy loss of the refrigerant while passing through themixing and pressurizing portion 16 e can be reduced, and the energy lossof the refrigerant when flowing out of the mixing and pressurizingportion 16 e can be reduced. FIG. 4 is a schematic diagram for onlyexplaining the sectional shape of the inner wall surface of the mixingand pressurizing portion 16 e, and the black points in FIG. 4 areindicated only for explaining the positions of the straight portion 16g, the curved line 101 and the curved line 102 in the section shape ofthe mixing and pressurizing portion 16 e.

As shown in FIG. 1, the first evaporator 17 is connected to thedownstream side of the mixing and pressurizing portion 16 e of theejector 16, that is, the outlet side of the expanding portion 16 h ofthe mixing and pressurizing portion 16 e. The first evaporator 17 is aheat exchanger, in which the refrigerant flowing out of the mixing andpressurizing portion 16 e of the ejector 16 is heat-exchanged with airblown by a blower fan 17 a, and is evaporated by absorbing heat from airpassing through the first evaporator 17.

The blower fan 17 a may be an electrical blower in which a fan rotationspeed is controlled by a control voltage output from an air conditioningcontroller (not shown) so as to control an air blowing amount. Arefrigerant outlet of the first evaporator 17 is coupled to therefrigerant suction port of the compressor 11.

In contrast, the second passage 14 b has the one end branched from thefirst passage 14 a at the branch portion 13, and the other end connectedto the refrigerant suction port 16 d of the ejector 16. A throttle unit18 and the second evaporator 19 are located in the second passage 14 bbetween the branch portion 13 and the refrigerant suction port 16 d ofthe ejector 16. The throttle unit 18 is configured to function as adecompression means for decompressing the refrigerant flowing into thesecond evaporator 19 via the second passage 14 b as well as a flowamount adjusting means for adjusting a flow amount of the refrigerantflowing into the second evaporator 19. As the throttle unit 18, a fixedthrottle such as a capillary tube an orifice or the like may be used, ora variable throttle may be used.

The second evaporator 19 is located in the second passage 14 b at adownstream side of the throttle unit 18, so that the refrigerantdecompressed in the throttle unit 18 flows into the second evaporator19. The second evaporator 19 is a heat exchanger, in which therefrigerant flowing out of the throttle unit 18 is heat-exchanged withair blown by a blower fan 19 a, and is evaporated by absorbing heat fromair passing through the second evaporator 19. The blower fan 19 a may bean electrical blower, similarly to the blower fan 17 a.

FIG. 5A shows a Mollier diagram showing refrigerant states in therefrigerant cycle of the refrigeration cycle device 10 with the abovestructure according to the first embodiment, and FIG. 5B is an enlargedview showing the part VB in FIG. 5A. When the compressor 11 is drivenand is operated by a power source such as a vehicle engine, thehigh-temperature and high-pressure refrigerant is discharged from thecompressor 100 (point 201 in FIG. 5A), and flows into the radiator 12.The high-temperature and high-pressure refrigerant is cooled andcondensed in the radiator 12 (from point 201 to point 202 in FIG. 5A).

The high-pressure refrigerant flowing out of the radiator 12 flows intothe receiver 12 b, and is separated into gas refrigerant and liquidrefrigerant. The separated liquid refrigerant flowing out of thereceiver 12 b flows into the branch portion 13 (from point 202 to point203 in FIG. 5A), and is branched into a refrigerant stream flowing intothe first passage 14 a so as to flow toward the nozzle 16 a and arefrigerant stream flowing into the second passage 14 b so as to flowtoward the refrigerant suction port 16 d.

A flow amount ratio Ge/Gnoz of the flow amount Ge of the refrigerantflowing through the second passage 14 b to the flow amount Gnoz of therefrigerant flowing through the first passage 14 a is determined basedon flow characteristics (decompression characteristics) of the expansionvalve 15, the nozzle 16 a of the ejector 16 and the throttle unit 18.

The refrigerant flowing into the expansion valve 15 through the branchedfirst passage 14 a is decompressed and expanded in the expansion valve15 while the flow amount of the refrigerant to flow into the nozzle 16 aof the ejector 16 is adjusted by the expansion valve 15 (from point 203to point 204 in FIG. 5A). Here, the flow amount of the refrigerant isadjusted by the expansion valve 15 so that the super-heat degree of therefrigerant at the refrigerant outlet side (the point 208 of FIG. 5A) ofthe first evaporator 17 is approached to a predetermined value. As shownin FIG. 5A from point 203 to point 204, the refrigerant is decompressedin iso-enthalpy in the expansion valve 15.

The refrigerant after being decompressed in the expansion valve 15 isfurther decompressed in the nozzle 16 a substantially in iso-entropywhile the enthalpy of the refrigerant is reduced (from point 204 topoint 205 in FIG. 5A). The pressure energy of the refrigerant isconverted to the speed energy of the refrigerant in the nozzle 16 a sothat the refrigerant is jetted from the refrigerant jet port 16 c of thenozzle 16 a by a high speed.

By the high-speed refrigerant stream from the refrigerant jet port 16 cof the nozzle 16 a, the refrigerant evaporated in the second evaporator19 is drawn into the ejector 16 from the refrigerant suction port 16 d.In FIG. 5A, ΔH indicates a reduction part of the enthalpy while therefrigerant is decompressed and expanded in iso-entropy at the nozzle 16a.

The refrigerant jetted from the nozzle 16 a and the refrigerant drawnfrom the refrigerant suction port 16 d flows into the mixing andpressurizing portion 16 e positioned downstream of the nozzle 16 a. Therefrigerant jetted from the nozzle 16 a and the refrigerant drawn fromthe refrigerant suction port 16 d are mixed in the mixing andpressurizing portion 16 e, and the speed energy of the refrigerant isconverted to the pressure energy, thereby increasing the refrigerantpressure in the mixing and pressurizing portion 16 e (point 205→to point206→point 207 in FIG. 5A).

The refrigerant flowing out of the mixing and pressurizing portion 16 eof the ejector 16 flows into the first evaporator 17. In the firstevaporator 17, low-pressure refrigerant is evaporated by absorbing heatfrom air blown by the blower fan 17 a, so that the enthalpy of therefrigerant is increased (from point 207 to point 208 in FIG. 5A). Thus,air passing through the first evaporator 17 is cooled and the cooled aircan be blown into a compartment to be cooled (e.g., a vehiclecompartment). The gas refrigerant flowing out of the first evaporator 17is drawn into the compressor 11 to be compressed again by the compressor11 (from point 208 to point 201 in FIG. 5A).

In contrast, the refrigerant stream flowing into the second passage 14 bfrom the branch portion 13 is decompressed and expanded by the throttleunit 18 (from point 203 to point 209 in FIG. 5A), and low-pressurerefrigerant decompressed by the throttle unit 18 flows into the secondevaporator 19. In the second evaporator 19, low-pressure refrigerant isevaporated by absorbing heat from air blown by the blower fan 19 a, sothat the enthalpy of the refrigerant is increased (from point 209 topoint 210 in FIG. 5A). Thus, air passing through the second evaporator19 is cooled and the cooled air can be blown into a compartment to becooled (e.g., the vehicle compartment).

The refrigerant after passing through the second evaporator 19 is drawninto the ejector 16 from the refrigerant suction port 16 d. Therefrigerant drawn from the refrigerant suction port 16 d flows into themixing and pressurizing portion 16 e of the ejector 16 through thesuction passage 16 i. In the present embodiment, the flow velocity ofthe refrigerant flowing through the suction passage 16 i is greater thanthe sound velocity, and the refrigerant passing through the suctionpassage 16 i is decompressed in iso-entropy as shown in FIG. 5B frompoint 210 to point 210′. While the refrigerant is decompressed in thesuction passage 16 i in iso-entropy, the enthalpy of the refrigerant isreduced by Δh.

The refrigerant flowing from the refrigerant suction port 16 d into themixing and pressurizing portion 16 e through the suction passage 16 i ismixed with the refrigerant jetted from the nozzle 16 a in the mixing andpressurizing portion 16 e. (from point 210′ to point 206 in FIG. 5A).Then, the refrigerant is pressurized in the mixing and pressurizingportion 16 e (from point 206 to point 207 in FIG. 5A), and is suppliedto the first evaporator 17 after passing through the mixing andpressurizing portion 16 e.

In the refrigeration cycle device 10 having the ejector 16, therefrigerant flowing out of the mixing and pressurizing portion 16 e ofthe ejector 16 can be supplied to the first evaporator 17 while therefrigerant decompressed by the throttle unit 18 in the second passage14 b can be supplied to the second evaporator 19 through the throttleunit 18. Thus, both the first evaporator 17 and the second evaporator 19can be operated simultaneously to have cooling functions.

Because a refrigerant downstream side of the first evaporator 17 isconnected to the refrigerant suction side of the compressor 11, therefrigerant pressurized in the mixing and pressurizing portion 16 e ofthe ejector 16 is drawn into the compressor 11. Therefore, the suctionpressure of the compressor 11 can be increased, and the driving power ofthe compressor 11 can be reduced. As a result, the coefficient ofperformance (COP) in the refrigerant cycle of the refrigeration cycledevice 10 can be effectively improved.

In the ejector 16 of the first embodiment, the suction passage 16 i isprovided to decompress the refrigerant in iso-entropy, such that theflow velocity of the refrigerant flowing into the mixing andpressurizing portion 16 e from the suction passage 16 i is equal to orlarger than the sound velocity. Therefore, the flow velocity of therefrigerant flowing into the mixing and pressurizing portion 16 e fromthe suction passage 16 i can be made substantially equal to the flowvelocity of the refrigerant jetted from the refrigerant jet port 16 c ofthe nozzle 16 a into the mixing and pressurizing portion 16 e.Therefore, the flow velocity of the refrigerant drawn from therefrigerant suction port 16 d can be increased while the energy loss ofthe refrigerant passing through the suction passage 16 i can be reduced.

Accordingly, the flow velocity of the gas refrigerant flowing into thestraight portion 16 g of the mixing and pressurizing portion 16 e can beincreased, and thereby the terminal velocity of grains of the liquidrefrigerant can be increased.

Thus, even when gas-liquid two-phase refrigerant passes through themixing and pressurizing portion 16 e in the ejector 16, the pressurizingamount of gas refrigerant can be increased in the mixing andpressurizing portion 16 e, thereby improving the ejector efficiency ηe.That is, even in the ejector 16 in which the kinetic energy ofgas-liquid two-phase refrigerant is converted to the pressure energythereof, the pressurizing amount of the gas refrigerant can beeffectively increased in the mixing and pressurizing portion 16 e.

In the present embodiment, because the refrigerant from the refrigerantsuction port 16 d is decompressed in the suction passage 16 i iniso-entropy as shown in FIG. 5B, the energy to be used for pressurizingcan be increased by the Δh, as compared with a case where therefrigerant is decompressed in iso-enthalpy. Thus, the pressurizingamount in the mixing and pressurizing portion 16 e can be increased byan amount corresponding to the Δh.

The ejector efficiency ηe′ of the present embodiment can be defined asin the following formula (F2) which is different from the formula (F1).

ηe′=((Gnoz+Ge)×(ΔP/ρ)/(Gnoz×Δi+Ge×Δh)   (F2)

Here, Ge is the flow amount of the suction refrigerant in the suctionpassage 16 i, Gnoz is the flow amount of the jet refrigerant jetted fromthe nozzle 16 a, ΔP is the pressurizing amount in the mixing andpressurizing portion 16 e, ρ is the density of the suction fluid, Δi isthe enthalpy difference between the inlet and the outlet of the nozzle16 a, and Δh is the energy to be used for pressurizing. As compared withthe above formula (F1), the expansion energy item (Ge×Δh) in the suctionpassage 16 i can be added in the denominator item (recovery energy item)in the formula (F2).

Thus, in the present embodiment, if the various configurations of theejector 16 are set such that the same ejector efficiency ηe in formulaF1 is obtained, the pressurizing amount ΔP can be increased by therecovery energy, thereby effectively improving the ejector efficiency.

In the present embodiment, the enthalpy reduction part Δh of therefrigerant while being decompressed and expanded in iso-entropy in thesuction passage 16 i, and the enthalpy reduction part ΔH of therefrigerant while being decompressed and expanded in iso-enthalpy in thenozzle 16 a have the following relationship in the formula F3.

ΔH≧Δh   (F3)

That is, in the present embodiment, respective configurations of theejector 16 are set to satisfy the above formula F3. That is, ΔH is theenthalpy difference between the enthalpy of the refrigerant at the inletof the nozzle 16 a and the enthalpy of the refrigerant at therefrigerant jet port 16 c of the nozzle 16 a, and Δh is the enthalpydifference between the enthalpy of the refrigerant at the inlet of thesuction passage 16 i and the enthalpy of the refrigerant at the outletof the suction passage 16 i.

According to the first embodiment of the present invention, because therefrigerant is decompressed and expanded in the suction passage 16 i iniso-entropy, the flow velocity of the refrigerant flowing into themixing and pressurizing portion 16 e from the suction passage 16 i canbe increased. If the iso-entropy decompression amount of the refrigerantin the suction passage 16 i is increased more than a necessary amount,the flow velocity of the refrigerant flowing from the suction passage 16i into the mixing and pressurizing portion 16 e is unnecessarilyincreased as compared with the flow velocity of the refrigerant jettedfrom the nozzle 16 a into the mixing and pressurizing portion 16 e.Therefore, the energy loss may be increased while the gas refrigerantand the liquid refrigerant having different flow velocities are mixed inthe mixing and pressurizing portion 16 e, and thereby the ejectorefficiency may be decreased.

That is, the unnecessary increased flow velocity of the refrigerant inthe suction passage 16 i causes the gas refrigerants having differentflow speeds to be mixed in the mixing and pressurizing portion 16 e,thereby increasing the energy loss and decreasing the ejectorefficiency.

FIG. 6 shows variations in the flow velocities of the gas refrigerantand liquid refrigerants within the ejector 16 when ΔH≧Δh and when ΔH<Δh.In the graphs of FIG. 6, the horizontal axis indicates axial positionsin the ejector 16 from the inlet of the nozzle 16 a to the outlet of theejector 16. The upper side graph of FIG. 6 indicates the presentembodiment where ΔH≧Δh, in which GJ2 indicates variations in the flowvelocity of gas refrigerant in the jet refrigerant jetted from thenozzle 16 a, GS2 indicates variations in the flow velocity of gasrefrigerant in the suction refrigerant drawn from the refrigerantsuction port 16 d, the chain line of LIQUID indicates variations in theflow velocity of liquid refrigerant. The lower side graph of FIG. 6indicates a comparison example where ΔH<Δh, in which GJ1 indicatesvariations in the flow velocity of gas refrigerant in the jetrefrigerant jetted from the nozzle 16 a, GS1 indicates variations in theflow velocity of gas refrigerant in the suction refrigerant drawn fromthe refrigerant suction port 16 d, the chain line of LIQUID indicatesvariations in the flow velocity of liquid refrigerant.

More specifically, the upper side graph in FIG. 6 shows the firstembodiment of the present invention in which the flow velocity GS2 ofthe gas refrigerant in the suction refrigerant flowing into the mixingand pressurizing portion 16 e from the suction passage 16 i isapproximately equal to the flow velocity GJ2 of gas refrigerant in thejet refrigerant flowing into the mixing and pressurizing portion 16 efrom the nozzle 16 a, and thereby ΔH≧Δh.

In contrast, the lower side graph in FIG. 6 shows the comparison examplein which the flow velocity GS1 of the gas refrigerant in the suctionrefrigerant flowing into the mixing and pressurizing portion 16 e fromthe suction passage 16 i is greatly faster than the flow velocity GJ1 ofgas refrigerant in the jet refrigerant flowing into the mixing andpressurizing portion 16 e from the nozzle 16 a, and thereby ΔH<Δh.

As shown in the graphs of FIG. 6, if the flow velocity of the gasrefrigerant in the suction refrigerant flowing into the mixing andpressurizing portion 16 e from the suction passage 16 i is greatlyfaster than the flow velocity of gas refrigerant in the jet refrigerantflowing into the mixing and pressurizing portion 16 e from the nozzle16, the flow of the gas refrigerant in the jet refrigerant isaccelerated by the flow of the gas refrigerant in the suctionrefrigerant. When the flow velocity of the gas refrigerant in the jetrefrigerant becomes equal to the flow velocity of the gas refrigerant inthe suction refrigerant, the gas refrigerant in the jet refrigerant andthe gas refrigerant in the suction refrigerant are joined with the sameflow velocity. Then, after the gas refrigerant in the jet refrigerantand the gas refrigerant in the suction refrigerant are joined with thesame flow velocity, the grains of the liquid refrigerant in the jetrefrigerant are accelerated by the joined gas refrigerant.

Accordingly, the terminal velocity, at which the flow velocity of theliquid refrigerant reaches to the joined flow velocity of the gasrefrigerant in the jet refrigerant and the gas refrigerant in thesuction refrigerant, is positioned on a downstream side in the mixingand pressurizing portion 16 e, and thereby the moving distance of theliquid refrigerant from the inlet to a position corresponding to theterminal velocity in the mixing and pressurizing portion 16 e isincreased. As a result, the distance from the position corresponding tothe terminal velocity to the outlet of the mixing and pressurizingportion 16 e, in which the kinetic energy is transmitted between the gasrefrigerant and the liquid refrigerant after the grains of the liquidrefrigerant reaches to the terminal velocity in the mixing andpressurizing portion 16 e, becomes shorter, and thereby the refrigerantcan not be sufficiently pressurized in the mixing and pressurizingportion 16 e.

In contrast, according to the first embodiment of the present invention,the configurations including dimensions in the respective portions ofthe ejector 16 are set so that ΔH≧Δh. Thus, it can prevent the flowvelocity of the suction refrigerant flowing from the suction passage 16i into the mixing and pressurizing portion 16 e from being excessivelyincreased.

More specifically, the tilt of the iso-entropy line of the gasrefrigerant from the inlet to the outlet of the suction passage 16 i(point 210 to point 210′ of FIGS. 5A, 5B) relative to a horizontal lineis smaller than the tilt of the iso-entropy line of the gas refrigerantfrom the inlet to the outlet of the nozzle (point 204 to point 205 ofFIG. 5A) relative to the horizontal line. Therefore, the decompressionamount of the refrigerant in the suction passage 16 i can be accuratelyset smaller than the decompression amount of the refrigerant in thenozzle 16 a. Thus, the refrigerant can be decompressed in the suctionpassage 16 i by a suitable decompression amount.

As a result, the energy loss, generated while gas refrigerants havingdifferent flow velocities are mixed, can be reduced, and the refrigerantcan be sufficiently pressurized in the mixing and pressurizing portion16 e, thereby effectively improving the ejector efficiency.

According to the first embodiment of the present invention, because thestraight portion 16 g is provided in a suitable range at a refrigerantinlet side of the mixing and pressurizing portion 16 e, the energy forceof the gas refrigerant can be effectively applied to the grains of theliquid refrigerant in the straight portion 16 g, and thereby the flowvelocity of the grains of the liquid refrigerant can rapidly reach tothe terminal velocity in the straight portion 16 g.

Furthermore, the kinetic energy of the liquid refrigerant having beingreached to the terminal velocity can be effectively transmitted to thegas refrigerant in the expanding portion 16 h. As a result, the energytransmission loss between the gas refrigerant and the liquid refrigerantin the expanding portion 16 h can be reduced, and thereby the ejectorefficiency can be sufficiently improved.

FIG. 7A shows the flow velocity of the gas refrigerant, the flowvelocity of the liquid refrigerant, the pressurizing amount ΔP shown bythe line P1, at respective positions from the inlet to the outlet of themixing and pressurizing portion 16 e of the ejector 16 according to thepresent embodiment. On the other hand, FIG. 7B shows the flow velocityof the gas refrigerant, the flow velocity of the liquid refrigerant, thepressurizing amount ΔP shown by the line P2, at respective positionsfrom the inlet of a mixing portion to the outlet of a diffuser portionof an ejector of a comparative example.

As shown in FIG. 7A, the straight portion 16 g is provided in a range ofthe mixing and pressurizing portion 16 e from the inlet of the mixingand pressurizing portion 16 e, such that the flow velocities of the gasrefrigerant and the liquid refrigerant in the refrigerant flowing intothe mixing and pressurizing portion 16 e become equal at the downstreamend of the straight portion 16 g. That is, the terminal velocity iscaused at the downstream end of the straight portion 16 g. Therefore,the kinetic energy of the refrigerant immediately after reaching to theterminal velocity can be converted to the pressure energy in theexpanding portion 16 h.

Because the flow velocity of the liquid refrigerant reaches to theterminal velocity at the inlet side of the expanding portion 16 h, theenergy transmission loss between the gas refrigerant and the liquidrefrigerant can be effectively reduced. Thus, the flow velocity of theliquid refrigerant and the gas refrigerant at the outlet of theexpanding portion 16 h can be sufficiently reduced, and the ratio ofenergy to be used actually for pressurizing can be increased.

As a result, the pressurizing amount ΔP of the refrigerant in the mixingand pressurizing portion 16 e can be increased in the present embodimentshown by the graph P1 in FIG. 7A, as compared with the comparativeexample shown by the graph P2 in FIGS. 7A and 7B.

According to experiments by the inventors of the present application,when the ratio L1/L2 of the length L1 of the straight portion 16 g tothe length L2 of the mixing and pressurizing portion 16 e from the inletto the outlet of the mixing and pressurizing portion 16 e is set aboutat 0.2, the pressurizing amount ΔP of the refrigerant in the mixing andpressurizing portion 16 e can be made maximum.

When the ratio L1/L2 is set about at 0.2, the flow velocities of the gasrefrigerant and the liquid refrigerant flowing out of the outlet of thestraight portion 16 g can be made approximately equal, and thepressurizing amount ΔP of the refrigerant in the mixing and pressurizingportion 16 e can be made maximum. If the manufacturing error of theejector 16 and the variation in the flow amount of the refrigerantcirculating in the refrigerant cycle of the refrigeration cycle device10 are considered, the ejector efficiency can be sufficiently increasedwhen 0<L1/L2≦0.4. More preferably, the ratio L1/L2 is set such that0.1<L1/L2≦0.3.

In a case where 0<L1/L2≦0.4, the ejector efficiency can be sufficientlyimproved even when the gas-liquid density difference of the gas-liquidtwo-phase refrigerant passing through the mixing and pressurizingportion 16 e is changed in a wider range of 0.9-600 kg/M³.

In the first embodiment of the present invention, the refrigerant can bepressurized substantially in iso-entropy in the entire range of themixing and pressurizing portion 16 e, and the sectional shape of themixing and pressurizing portion 16 e is changed so as to reduce aseparation from of the refrigerant at the outlet of the mixing andpressurizing portion 16 e. Therefore, the energy loss of the refrigerantpassing through the mixing and pressurizing portion 16 e can be reduced,thereby reducing energy loss of the refrigerant flowing out of themixing and pressurizing portion 16 e.

As a result, the ratio of the energy to be actually used for pressuringcan be increased among the recovery energy in the ejector 16. FIG. 8shows energy distribution in the mixing and pressurizing portion 16 e ofthe ejector 16 according to the first embodiment and the comparisonexample. In FIG. 8, E1 indicates the energy to be used for pressurizing,E2 indicates the remaining energy of the refrigerant, E3 indicates theenergy transmission loss, and E4 indicates the other loss. As shown inFIG. 8, according to the first embodiment, the energy to be used forpressurizing in the mixing and pressurizing portion 16 e can be greatlyincreased as compared with the comparison example.

Second Embodiment

A second embodiment of the present invention will be described withreference to FIG. 9. FIG. 9 is a diagram corresponding to FIG. 3 of theabove-described first embodiment. In the second embodiment, the passagesectional area of the suction passage 16 i is changed such that theratio (passage area ratio) of the passage sectional area of the suctionpassage 16 i to the passage sectional area at the inlet of the suctionpassage 16 i is changed as in the straight line graph shown in FIG. 9.As shown in FIG. 9, the passage sectional area of the suction passage 16i is changed from the inlet to the outlet of the suction passage 16 i bya constant degree. In the second embodiment, the other parts of theejector 16 are similar to those in the ejector 16 of the above-describedfirst embodiment.

According to the second embodiment of the present invention, the suctionpassage 16 i of the ejector 16 can be configured, such that the flowvelocity of the suction refrigerant flowing from the suction passage 16i into the straight portion 16 g of the mixing and pressurizing portion16 e becomes equal to or greater than the sound velocity and the suctionrefrigerant is decompressed in iso-entropy. Thus, the terminal velocityof the grains of the liquid refrigerant flowing into the straightportion 16 g in the mixing and pressurizing portion 16 e can beincreased, thereby improving the ejector efficiency. In the secondembodiment, the other parts of the ejector 16 are similar to those inthe ejector 16 of the above-described first embodiment.

Third Embodiment

A third embodiment of the present invention will be described withreference to FIG. 10. FIG. 10 is a diagram corresponding to FIG. 3 ofthe above-described first embodiment. In the third embodiment, thepassage sectional area of the suction passage 16 i is changed, such thatthe passage sectional area at the inlet side of the suction passage 16 iis gradually reduced toward downstream in the refrigerant flow directionfrom the inlet of the suction passage 16 i, and the passage sectionalarea at the outlet side of the suction passage 16 i is graduallyincreased toward downstream in the refrigerant flow direction. That is,at a predetermined portion between the inlet and the outlet of thesuction passage 16 i, the passage sectional area of the suction passage16 i becomes smallest, as shown in FIG. 10. A reduction ratio of thepassage sectional area at the inlet side of the suction passage 16 i islarger than an increase ratio of the passage sectional area at theoutlet side of the suction passage 16 i. At the outlet side of thesuction passage 16 i, the passage sectional area of the suction passage16 i is gradually increased, but is not increased more than the passagesectional area at the inlet of the suction passage 16 i.

In the third embodiment, the other parts of the ejector 16 are similarto those in the ejector 16 of the above-described first embodiment.

According to the third embodiment of the present invention, the suctionpassage 16 i of the ejector 16 is configured such that the flow velocityof the suction refrigerant flowing through the suction passage 16 ibecomes equal to or greater than the sound velocity at a contractionposition where the refrigerant passage area becomes smallest in thesuction passage 16 i. Thus, the flow velocity of the suction refrigerantcan be increased downstream of the contraction position in the suctionpassage 16 i. Therefore, the terminal velocity of the grains of theliquid refrigerant flowing into the straight portion 16 g in the mixingand pressurizing portion 16 e can be increased, thereby improving theejector efficiency.

Fourth Embodiment

A fourth embodiment of the present invention will be described withreference to FIG. 11. FIG. 11 is a schematic diagram corresponding toFIG. 4 of the above-described first embodiment, showing a refrigerantpassage sectional shape of the expending portion 16 h in a sectionincluding the center axis of the nozzle 16 a of the ejector 16. As shownin FIG. 11, the passage wall surface of the expending portion 16 h isconfigured by combining plural straight line portions 103, 104, 105,106, 107. That is, the expanding portion 16 h is formed by pluralcylindrical passage portions (103 to 107) each of which has a tapersurface. The taper surfaces of the plural cylindrical passage portions(103 to 107) are suitably combined so as to form the expending portion16 h in the mixing and pressurizing portion 16 e.

In the fourth embodiment, the other parts of the ejector 16 are similarto those in the ejector 16 of the above-described first embodiment.

In the structure of the expanding portion 16 h according to the fourthembodiment, the energy transmission loss between the gas refrigerant andthe liquid refrigerant can be reduced, thereby sufficiently improvingthe ejector efficiency. In the above example of the fourth embodiment,the structure of the expanding portion 16 h is used for the ejector 16according to the first embodiment. However, the structure of theexpanding portion 16 h of the fourth embodiment can be used for theejector 16 according to any one of the second and third embodiments ofthe present invention.

Fifth Embodiment

A fifth embodiment of the present invention will be described withreference to FIG. 12. FIG. 12 is a schematic diagram corresponding toFIG. 4 of the above-described first embodiment, showing a refrigerantpassage sectional shape of the expending portion 16 h in a sectionincluding the center axis of the nozzle 16 a of the ejector 16. As shownin FIG. 12, the passage wall surface of the expending portion 16 h isconfigured by combining plural straight line portions 103, 104, 105, andthe curved line portion 102. That is, the expanding portion 16 h isformed, by plural cylindrical passage portions (103 to 105) each ofwhich has a taper surface and by the cylindrical passage portion (102)having a curved surface (102).

In the fifth embodiment, the other parts of the ejector 16 are similarto those in the ejector 16 of the above-described first embodiment.

In the structure of the expanding portion 16 h according to the fifthembodiment, the energy transmission loss between the gas refrigerant andthe liquid refrigerant can be reduced, thereby sufficiently improvingthe ejector efficiency. In the above example of the fifth embodiment,the structure of the expanding portion 16 h is used for the ejector 16according to the first embodiment. However, the structure of theexpanding portion 16 h of the fifth embodiment can be used for theejector 16 according to any one of the second and third embodiments ofthe present invention.

Sixth to Eighth Embodiments

A sixth embodiment of the present invention will be described withreference to FIG. 13. FIG. 13 is a schematic diagram corresponding toFIG. 4 of the above-described first embodiment, showing a refrigerantpassage sectional shape of the expending portion 16 h in a sectionincluding the center axis of the nozzle 16 a of the ejector 16. As shownin FIG. 13, the passage wall surface of the expending portion 16 h isconfigured by a single straight line portion 108 having a constant taperangle. That is, the refrigerant passage sectional area of the expandingportion 16 h in the mixing and pressurizing portion 16 e is graduallyincreased toward downstream by a constant expanding degree in the entirelength of the expanding portion 16 h.

A seventh embodiment of the present invention will be described withreference to FIG. 14. FIG. 14 is a schematic diagram corresponding toFIG. 4 of the above-described first embodiment, showing a refrigerantpassage sectional shape of the expending portion 16 h in a sectionincluding the center axis of the nozzle 16 a of the ejector 16. As shownin FIG. 14, the passage wall surface of the expending portion 16 h isconfigured by combining plural straight line portions 103, 104, 105,106, 109. The plural straight line portions 103, 104, 105, 106, 109 aresuitably combined to configure the expanding portion 16 h such that theexpanding degree of the refrigerant passage sectional area of theexpanding portion 16 h is gradually increased.

An eighth embodiment of the present invention will be described withreference to FIG. 15. FIG. 15 is a schematic diagram corresponding toFIG. 4 of the above-described first embodiment, showing a refrigerantpassage sectional shape of the expending portion 16 h in a sectionincluding the center axis of the nozzle 16 a of the ejector 16. As shownin FIG. 15, the passage wall surface of the expending portion 16 h isconfigured by a single curved line portion 110 in which the expendingangle is gradually increased as toward downstream.

In the sixth to eighth embodiments of the present invention, the otherparts of the ejector 16 can be made similar to those of the ejector 16according to the first embodiment, and the ejector efficiency can beincreased. In the above examples of the sixth to eighth embodiments ofthe present invention, the structure of the expanding portion 16 h isused for the ejector 16 according to the first embodiment. However, thestructure of the expanding portion 16 h according to any one of thesixth to eighth embodiments can be used for the ejector 16 of any one ofthe second and third embodiments of the present invention. That is, theexpanding portion 16 h of the mixing and pressurizing portion 16 eaccording to any one of the fourth to eighth embodiments can be suitablyused for the ejector 16 according to any one of the first to thirdembodiments.

Ninth Embodiment

In the above-described embodiments of the present invention, the ejector16 includes the mixing and pressurizing portion 16 e that is configuredby the straight portion 16 g and the expanding portion 16 h. However, ina ninth embodiment of the present invention, an ejector 16 is configuredwithout using the straight portion 16 g, so that the mixing andpressurizing portion 16 e is configured only by the expanding portion 16h as shown in FIG. 16.

FIG. 16 is an axial sectional view of the ejector 16 according to theninth embodiment of the present invention, which corresponds to FIG. 2Aof the first embodiment. That is, the inlet of the expanding portion 16h is located at the position corresponding to the refrigerant jet port16 c of the nozzle 16 a in the axial direction of the nozzle 16 a. Thepassage wall surface of the expanding portion 16 h of the ejector 16shown in FIG. 16 has the same passage sectional shape of the expendingportion 16 h in the mixing and pressurizing portion 16 e of theabove-described first embodiment shown in FIG. 4. Thus, the innerperipheral surface of the expending portion 16 h is curved to be convextoward radially inside at the inlet side of the expanding portion 16 h,and is curved to be convex toward radially outside at the outlet side ofthe expanding portion 16 h.

Thus, even when the straight portion 16 g is omitted in the mixing andpressurizing portion 16 e, it is possible for the inlet side of themixing and pressurizing portion 16 e to have the same function as thestraight portion 16 g, thereby improving the ejector efficiency. Themixing and pressurizing portion 16 e configured by only the expandingportion 16 h according to the ninth embodiment can be used for theejector 16 according to the second or the third embodiment.

Furthermore, in a case where the ejector efficiency can be sufficientlyincreased by increasing the flow velocity of the suction refrigerant inthe suction passage 16 i, the straight portion 16 g described in any oneof the above embodiments may be omitted from the mixing and pressurizingportion 16 e.

For example, in the above-described fourth to eighth embodiments shownin FIGS. 11 to 15, the mixing and pressurizing portion 16 e may beconfigured by only the expending portion 16 h without using the straightportion 16 g. In this case, the inlet of the expanding portion 16 h islocated at a position corresponding to the refrigerant jet port 16 c ofthe nozzle 16 a in the ejector 16 according to any one of the fourth toeighth embodiments.

Tenth Embodiment

In the ejector 16 according to any one of the above-describedembodiments, the suction passage 16 i is provided between the outerperipheral surface of the tip end portion of the nozzle 16 a and theinner peripheral surface of body portion 16 b. In the ejector 16 of thetenth embodiment, the nozzle 16 a is used as a first nozzle 16 a, and asecond nozzle 16 j is provided for forming a suction passage 16 ithrough which the refrigerant drawn from a refrigerant suction port 16 dflows into the mixing and pressurizing portion 16 e, as shown in FIG.17. That is, the suction passage 16 i is defined by the second nozzle 16j and the refrigerant suction port 16 d is provided at the inlet of thesecond nozzle 16 j, so that the refrigerant drawn from the suction port16 d flows into the mixing and pressurizing portion 16 e through thesuction passage 16 i.

As an example of the second nozzle 16 j in the tenth embodiment, a Lavalnozzle may be used. The refrigerant passage sectional area of thesuction passage 16 i of the second nozzle 16 j can be changed similar tothat of the suction passage 16 i of the third embodiment. In this case,the advantages of the suction passage 16 i described in the thirdembodiment can be obtained.

Alternatively, the second nozzle 16 j can be configured by a tapernozzle such that the refrigerant passage sectional area of the suctionpassage 16 i of the second nozzle 16 j is changed similar to that of thesuction passage 16 i of the first or second embodiment described above.In this case, the advantages of the suction passage 16 i described inthe first or second embodiment can be obtained.

Eleventh Embodiment

In the above-described embodiments, the ejector 16 is typically used forthe refrigeration cycle device 10 that is provided with the radiator 12and the receiver 12 b, for example, as shown in FIG. 1. In therefrigeration cycle device 10, the radiator 12 provided with thereceiver 12 b is an example of a super-cooled type condenser in whichthe refrigerant is cooled and condensed.

In the eleventh embodiment, the ejector 16 according to any one of theabove-described embodiments is used for a refrigeration cycle devicehaving a super-cooled type condenser that is configured by acondensation heat exchanging portion, a receiver portion and asuper-cooling heat exchanging portion. Here, the condensation heatexchanging portion is configured to cool and condense the high-pressurerefrigerant from the compressor 11, the receiver portion is configuredto separate the refrigerant flowing from the condensation heatexchanging portion into gas refrigerant and liquid refrigerant, and thesuper-cooling heat exchanging portion is configured to super-cool thesaturated liquid refrigerant from the receiver portion. Even in thiscase, the liquid refrigerant super-cooled in the super-cooling heatexchanging portion can be introduced into the branch portion 13 to bebranched at the branch portion 13. The other parts of the refrigerantcycle structure in the refrigeration cycle device of the eleventhembodiment may be similar to those of the refrigeration cycle device 10shown in FIG. 1.

FIG. 18 is a Mollier diagram showing refrigerant states in a refrigerantcycle of the refrigeration cycle device according to the eleventhembodiment, in which the super-cooled type condenser configured by thecondensation heat exchanging portion, the receiver portion and thesuper-cooling heat exchanging portion is used instead of the receiver 12provided with the receiver 12 b. In this case, as shown in FIG. 18, theliquid refrigerant of a super-cooled state (point 203′ of FIG. 18) isbranched at the branch portion 13.

Thus, the refrigerant state flowing from the expansion valve 15 into thenozzle 16 a of the ejector 16 may become in a gas-liquid two-phase state(point 204 of FIG. 18) or in a liquid state (point 204′ of FIG. 18). InFIG. 18, the parts corresponding to or similar to those in FIG. 5A areindicated by the same reference numbers, and the detail explanationthereof is omitted.

Even in the refrigeration cycle device with the Mollier diagram shown inFIG. 18, the ejector 16 is configured such that, the flow velocity ofthe suction refrigerant passing through the suction passage 16 i of theejector 16 is increased, the flow velocity of the grains of the liquidrefrigerant can be rapidly reached to the terminal velocity by thestraight portion 16 g of the mixing and pressurizing portion 16 e, andthe flow velocity of the refrigerant can be sufficiently reduced in theexpanding portion 16 h. Thus, the ejector efficiency can be improved.

Thus, even in a case where gas-liquid two-phase refrigerant flows intothe nozzle 16 a or only the liquid refrigerant flows into the nozzle 16a in the refrigerant cycle, when the mixed refrigerant, in which the jetrefrigerant jetted from the nozzle 16 a and the suction refrigerantflowing from the suction passage 16 i are mixed, is in the gas-liquidtwo-phase state in the ejector 16, the ejector efficiency can beeffectively improved.

Twelfth Embodiment

In the above-described embodiments, the ejector 16 is used for therefrigerant cycle in which the expansion valve 15 is provided in thefirst passage 14 a at an upstream side of the nozzle 16 a of the ejector16, for example, as shown in FIG. 1. However, in the twelfth embodiment,the ejector 16 is used for a refrigerant cycle of a refrigeration cycledevice in which the expansion valve 15 is omitted from the refrigerationcycle device 10 shown in FIG. 1. The other parts of the refrigerationcycle device of the twelfth embodiment are similar to those of therefrigeration cycle device 10 shown in FIG. 1. The refrigerant state ofthe refrigerant cycle is changed as in the Mollier diagram of FIG. 19when the refrigeration cycle device according to the twelfth embodimentis operated.

Because the expansion valve 15 is not provided in the refrigerationcycle device shown in FIG. 1, the refrigerant branched at the branchportion 13 flows into the nozzle 16 a of the ejector 16 through thefirst passage 14 a, and is decompressed and expanded substantially iniso-entropy in the nozzle 16 a (from point 203 to point 205 of FIG. 19).Even when the ejector 16 of the present invention is used for therefrigeration cycle device in which the refrigerant from the branchportion 13 without being decompressed is firstly decompressed andexpanded in the nozzle 16 a of the ejector 16, the ejector efficiencycan be improved similarly to that in the above-described firstembodiment.

Alternatively, both the receiver 12 b and the expansion valve 15 can beomitted from the refrigeration cycle device 10 shown in FIG. 1. In thiscase, gas-liquid two-phase refrigerant flowing out of the radiator 12 isdirectly branched at the branch portion 13 (point 202 of FIG. 19), andflows into the nozzle 16 a of the ejector 16 through the first passage14 a to be decompressed and expanded substantially in iso-entropy in thenozzle 16 a of the ejector 16. Alternatively, a super-cooled typecondenser can be used as the radiator 12 similarly to the abovedescribed in the eleventh embodiment while the expansion valve 15 isomitted in the refrigeration cycle device 10 shown in FIG. 1. In thiscase, a super-cooled liquid refrigerant (point 203′ of FIG. 19) flowingout of the radiator 12 is branched at the branch portion 13, and a partof the branched refrigerant flows into the nozzle 16 a of the ejector 16through the first passage 14 a to be decompressed substantially iniso-entropy in the nozzle 16 a.

Thirteenth Embodiment

In the above-described embodiments, the ejector 16 is used for arefrigerant cycle in which the refrigerant state flowing from the branchportion 13 into the first passage 14 a and the refrigerant state flowingfrom the branch portion 13 into the second passage 14 b are made equal.However, the ejector 16 can be used for a refrigerant cycle in which therefrigerant state flowing from the branch portion 13 into the firstpassage 14 a and the refrigerant state flowing from the branch portion13 into the second passage 14 b are made different from each other.

As an example of a refrigeration cycle device of a thirteenthembodiment, the receiver 12 b shown in FIG. 1 is omitted, the expansionvalve 15 is located upstream of the branch portion 13, and the branchportion 13 is configured so as to change the refrigerant states (e.g.,dryness) flowing into the first and second passages 14 a, 14 b.

For example, the branch portion 13 may be configured to have an interiorspace in which a scroll flow of the refrigerant is generated so that thedryness distributions of the refrigerant are caused in the interiorspace of the branch portion 13 by centrifugal force due to the scrollflow of the refrigerant.

The first passage 14 a and the second passage 14 b are connected to thebranch portion 13 so that refrigerant having a predetermined dryness canbe respectively introduced into the first passage 14 a and the secondpassage 14 b. Thus, the dryness of the refrigerant flowing into thefirst passage 14 a from the branch portion 13 and the dryness of therefrigerant flowing into the second passage 14 b from the branch portion13 can be suitably changed. As the structure of the branch portion 13,the structure described in US 2007/028630 (corresponding to JP2007-46806) can be incorporated herein by reference.

When the refrigeration cycle device according to the thirteenthembodiment is operated, refrigerant states circulated in the refrigerantcycle can be set to be changed as in the Mollier diagram shown in FIG.20 or FIG. 21. In the diagram of FIG. 20, the refrigerant flowing intothe nozzle 16 a of the ejector 16 from the branch portion 13 through thefirst passage 14 a is in a gas-liquid two-phase state (point 203″ inFIG. 20). On the other hand, in the diagram of FIG. 21, the refrigerantflowing into the nozzle 16 a of the ejector 16 from the branch portion13 through the first passage 14 a is in a liquid state (point 203′ inFIG. 21).

Even in the refrigerant cycle with the operation states shown in FIG. 20or FIG. 21, the ejector efficiency can be effectively improved by usingthe ejector 16 according to any one of the first to tenth embodiments.

Fourteenth Embodiment

In the above-described embodiments, the ejector 16 of the presentinvention is used for a sub-critical refrigerant cycle in which thepressure of refrigerant on a high-pressure side before beingdecompressed is lower than the critical pressure of the refrigerant.However, in a fourteenth embodiment of the present invention, theejector 16 is used for a super-critical refrigerant cycle in which thepressure of refrigerant on the high-pressure side before beingdecompressed is higher than the critical pressure of the refrigerant.For example, carbon dioxide is used as the refrigerant so that therefrigerant pressure discharged from the compressor 11 becomes higherthan the critical pressure of the refrigerant.

FIG. 22 shows an example of a refrigeration cycle device 10 according tothe fourteenth embodiment of the present invention. In the refrigerationcycle device 10 of FIG. 22, the receiver 12 b and the expansion valve 15are omitted from the refrigeration cycle device 10 shown in FIG. 1, anda pressure control valve is used as the throttle unit 18 as comparedwith the refrigerant cycle device 10 shown in FIG. 1. A valve opendegree of the throttle unit 18 is adjusted such that the refrigerantpressure on the high-pressure side of the refrigerant cycle of therefrigeration cycle device 10 is approached to a target pressure that isdetermined in accordance with a temperature of the refrigerant at arefrigerant outlet side of the radiator 12.

For example, the throttle unit 18 is provided with a temperature sensingportion 18 a located at the refrigerant outlet side of the radiator 12.The temperature sensing portion 18 a is configured to generate thereinan inner pressure corresponding to the temperature of the refrigerant onthe refrigerant outlet side of the radiator 12, so that the valve opendegree of the throttle unit 18 is adjusted by a balance between theinner pressure of the temperature sensing portion 18 a and the pressureof the refrigerant on the refrigerant outlet side of the radiator 12.Thus, the refrigerant pressure on the high-pressure side of therefrigerant cycle can be adjusted to the target pressure, and therebythe COP of the refrigerant cycle can be made maximum.

In the fourteenth embodiment, as shown in FIG. 22, an accumulator 20 asa low-pressure side gas-liquid separator is located at a refrigerantoutlet side of the first evaporator 17 so that surplus refrigerant inthe refrigerant cycle is stored in the accumulator 20. A gas refrigerantoutlet is provided in the accumulator 20 and is coupled to therefrigerant suction side of the compressor 11 so that the gasrefrigerant separated from the liquid refrigerant in the accumulator 20is supplied to the compressor 11. In the components of the refrigerationcycle device 10 shown in FIG. 22, the other parts are similar to thoseof the refrigeration cycle device 10 shown in FIG. 1.

When the refrigeration cycle device 10 of the present embodiment isoperated, the refrigerant state is charged as in the Mollier diagramshown in FIG. 23. As shown in FIG. 23, the refrigerant is compressed inthe compressor 11 to have a pressure higher than the critical pressureof the refrigerant (point 201 of FIG. 23), and is discharged to theradiator 12.

The refrigerant is cooled in the radiator 12 by performing heat exchangewith outside air while keeping the refrigerant pressure at the pressurehigher than the critical pressure (from point 201 to point 202 of FIG.23). The high-pressure refrigerant flowing out of the radiator 12 isbranched at the branch portion 13 into a refrigerant stream flowing intothe first passage 14 a and a refrigerant stream flowing into the secondpassage 14 b.

The refrigerant flowing into the first passage 14 a from the branchpassage 13 flows through the nozzle 16 a, the first evaporator 17 andthe accumulator 20 in this order (point 202→point 205→point 206→point207→point 208 in FIG. 23). The gas refrigerant separated at theaccumulator 20 is drawn into the compressor 11.

On the other hand, the refrigerant flowing into the second passage 14 bflows through the throttle unit 18 (i.e., high-pressure control valve)and the second evaporator 19 in this order, and is drawn into theejector 16 from the refrigerant suction port 16 d (point 202→point209→point 210→point 210′→point 206 in FIG. 23). The throttle unit 18 isadjusted so as to adjust the refrigerant pressure on the high pressureside from the refrigerant discharge side of the compressor 11 to theinlet of the nozzle 16 a of the ejector 16 and the inlet of the throttleunit 18, such that the COP of the refrigerant cycle becomes the targetpressure.

Thus, even in the refrigerant cycle of the refrigeration cycle device 10in which the super-critical refrigerant flows into the nozzle 16 a ofthe ejector 16, the ejector efficiency can be improved.

Even in a case where the super-critical refrigerant flows into thenozzle 16 a of the ejector 16, when the mixed refrigerant, in which thejet refrigerant jetted from the nozzle 16 a and the suction refrigerantdrawn from the refrigerant suction port 16 d are mixed, is in agas-liquid two-phase state in the ejector 16, the ejector efficiency canbe significantly improved.

That is, when the ejector 16 is used for a super-critical refrigerantcycle in which at least the jet refrigerant jetted from nozzle 16 d isin a gas-liquid two-phase state or the refrigerant downstream of thethroat portion of the nozzle 16 d is in a gas-liquid two-phase state,the ejector efficiency can be more significantly improved.

Fifteenth Embodiment

A fifteenth embodiment of the present invention will be described withreference to FIGS. 24 and 25. As shown in FIG. 24, in a refrigerationcycle device 10 of the fifteenth embodiment, the compressor 11 is usedas a first compressor 11, and a second compressor 21 is added in thesecond passage 14 b between the refrigerant outlet of the secondevaporator 19 and the refrigerant suction port 16 d of the ejector 16.Therefore, the second compressor 21 compresses the refrigerant flowingout of the second evaporator 19 and discharges the compressedrefrigerant to the refrigerant suction port 16 d of the ejector 16. Theother components of the refrigeration cycle device 10 of the fifteenthembodiment are similar to those of the refrigeration cycle device 10shown in FIG. 1.

For example, in the fifteenth embodiment of the present invention, thefirst evaporator 17 can be used for cooling the interior of a passengercompartment of a vehicle, and the second evaporator 19 can be used forcooling a cooler box (refrigerator) mounted in the vehicle. That is, thespace to be cooled by the first evaporator 17 is the passengercompartment of the vehicle, and the space to be cooled by the secondevaporator 19 is the interior space of the cooler box.

The basic structure of the second compressor 21 may be similar to thatof the first compressor 11, and a generally known compressor may be usedas the second compressor 21.

FIG. 25 is a Mollier diagram showing the refrigerant operation state ofthe refrigerant cycle of the refrigeration cycle device 10, according tothe fifteenth embodiment. As shown in FIG. 25, the refrigerant iscompressed in the first compressor 11 to be in a high-pressure andhigh-temperature state (point 201 of FIG. 25), and is discharged to theradiator 12. The high-pressure and high-temperature refrigerant iscooled in the radiator 12 by performing heat exchange with outside air(from point 201 to point 202 of FIG. 25). The high-pressure refrigerantflowing out of the radiator 12 is separated into gas refrigerant andliquid refrigerant in the receiver 12 b, and the separated liquidrefrigerant flows into the branch portion 13 (from point 202 to point203 of FIG. 25), similarly to FIG. 5A of the first embodiment. Then, therefrigerant is branched at the branch portion 13 into a refrigerantstream flowing into the first passage 14 a and a refrigerant streamflowing into the second passage 14 b.

The refrigerant flowing into the expansion valve 15 through the branchedfirst passage 14 a is decompressed and expanded in iso-enthalpy by theexpansion valve 15 (from point 203 to point 204 in FIG. 25). Then, therefrigerant after being decompressed at the expansion valve 15 isfurther decompressed and expanded in the nozzle 16 a substantially iniso-entropy while the enthalpy of the refrigerant is reduced (from point204 to point 205 in FIG. 25). The pressure energy of the refrigerant isconverted to the speed energy of the refrigerant in the nozzle 16 a sothat the refrigerant is jetted from the refrigerant jet port 16 c by ahigh speed. Then, the refrigerant jetted from the refrigerant jet port16 c of the nozzle 16 is mixed in the mixing and pressurizing portion 16e with the refrigerant drawn from the refrigerant suction port 16 d, sothat the mixed refrigerant is pressurized in the mixing and pressurizingportion 16 e (from point 206 to point 207 in FIG. 25).

The refrigerant flowing out of the mixing and pressurizing portion 16 eof the ejector 16 flows into the first evaporator 17. In the firstevaporator 17, low-pressure refrigerant is evaporated by absorbing heatfrom air blown by the blower fan 17 a, so that the enthalpy of therefrigerant is increased (from point 207 to point 208 in FIG. 25). Thus,air passing through the first evaporator 17 is cooled and the cooled aircan be blown into the passenger compartment. The gas refrigerant flowingout of the first evaporator 17 is drawn into the first compressor 11 tobe compressed again by the first compressor 11 (from point 208 to point201 in FIG. 25).

In contrast, the refrigerant stream flowing into the second passage 14 bfrom the branch portion 13 is decompressed and expanded in iso-enthalpyby the throttle unit 18 (from point 203 to point 209 in FIG. 25), andlow-pressure refrigerant decompressed by the throttle unit 18 flows intothe second evaporator 19. In the second evaporator 19, low-pressurerefrigerant is evaporated by absorbing heat from air blown by the blowerfan 19 a, so that the enthalpy of the refrigerant is increased (frompoint 209 to point 210 in FIG. 25). Thus, air passing through the secondevaporator 19 is cooled so as to cool the interior of the cooler box.

In the fifteenth embodiment of the present invention, the throttledpassage area of the throttle unit 18 can be set smaller than that of thefirst embodiment, thereby increasing the refrigerant decompressionamount at the throttle unit 18. Therefore, the refrigerant evaporationpressure (refrigerant evaporation temperature) in the second evaporator19 can be set lower as compared with the first embodiment.

As shown in FIG. 24, the refrigerant flowing out of the secondevaporator 19 is drawn into the second compressor 21, and is compressedin the second compressor 21 (from point 210 to point 211 in FIG. 25).Then, the compressed refrigerant is discharged from the secondcompressor 21 into the refrigerant suction port 16 d of the ejector 16,and is drawn into the mixing and pressurizing portion 16 e of theejector 16 from the refrigerant suction port 16 d. Similarly to thefirst embodiment, the refrigerant is decompressed in iso-entropy whilepassing through the suction passage 16 i (from point 211 to point 210′in FIG. 25). The other operations of the refrigeration cycle device 10are similar to those of the above-described first embodiment.

In the refrigeration cycle device 10 having the ejector 16 according tothe fifteenth embodiment, the refrigerant flowing out of the mixing andpressurizing portion 16 e of the ejector 16 can be supplied to the firstevaporator 17 while the refrigerant greatly decompressed by the throttleunit 18 in the second passage 14 b can be supplied to the secondevaporator 19 through the throttle unit 18. Thus, both the firstevaporator 17 and the second evaporator 19 can be operatedsimultaneously to have greatly different cooling capacities, and therebythe second evaporator 19 can be used to cool the interior of the coolerbox that needs a cooling temperature lower than that in the passengercompartment.

At a low outside air temperature, a pressure difference between therefrigerant pressure on the high-pressure side and the refrigerantpressure on the low-pressure side becomes smaller in the refrigerantcycle of the refrigerant cycle device 10. In this case, the flow amountof the refrigerant passing through the nozzle 16 a of the ejector 16 maybe decreased, and thereby the suction capacity of the ejector 13 may bedecreased. Even in this case, because the second compressor 21 islocated in the refrigeration cycle device 10 of the fifteenthembodiment, the suction capacity of the refrigerant into the ejector 16from the refrigerant suction port 16 d can be increased, so that therefrigerant cycle can be stably operated.

Furthermore, because the refrigerant is pressurized by using both thefirst and second compressors 11, 21, a pressure difference between thesuction pressure and the discharge pressure in respective compressors11, 21 can be reduced. Thus, compression efficiency of each of the firstand second compressors 11, 21 can be improved, thereby improving the COPin the refrigerant cycle of the refrigeration cycle device 10.

The compression efficiency in the compressor 11, 21 is a ratio ΔE1/ΔE2of an increase amount ΔE1 of the enthalpy of the refrigerant while beingcompressed in iso-entropy in the compressor 11, 21 to an increase amountΔE2 of the enthalpy of the refrigerant while being actually compressedin the compressor 11, 21. For example, when the rotational speed or thepressurizing amount of the compressor 11, 21 is increased, therefrigerant temperature is increased by the friction force, and therebythe increase amount ΔE2 is increased and the compression efficiency isreduced.

Thus, in the refrigeration cycle device 10, if the pressure differencebetween the refrigerant pressure on the high-pressure side and therefrigerant pressure on the low-pressure side needs to be increased, theimproving effect of the COP in the refrigerant cycle can be madesignificantly.

According to the fifteenth embodiment of the present invention, evenwhen the ejector 16 is used for the refrigerant cycle device 10 providedwith the first compressor 11 and the second compressor 21, the ejectorefficiency can be sufficiently improved. Furthermore, the refrigerantsuction capacity of the ejector 16 can be suitably increased by usingthe second compressor 21, and thereby the configuration of the ejector16 can be easily set.

Thus, in the present embodiment, the ejector 16 can be easily configuredso as to prevent the flow velocity of the suction refrigerant flowingfrom the suction passage 16 i into the mixing and pressurizing portion16 e from being unnecessarily increased. That is, in the presentembodiment, because the flow velocity of the refrigerant flowing fromthe suction passage 16 i into the mixing and pressurizing portion 16 ecan be changed by not only the decompression characteristics in thesuction passage 16 i but also the discharge refrigerant pressure of thesecond compressor 21, the suction passage 16 i of the ejector 16 can beeasily formed. Therefore, the flow velocity of the suction refrigerantflowing from the suction passage 16 i into the mixing and pressurizingportion 16 e can be easily adjusted by adjusting the refrigerantpressure at the refrigerant suction port 16 d of the ejector 16.

According to the fifteenth embodiment, by adjusting the refrigerantdischarge capacity of the second compressor 21, the flow velocity of thesuction refrigerant flowing from the suction passage 16 i into themixing and pressurizing portion 16 e can be easily adjusted at asuitable velocity, relative to the flow velocity of the jet refrigerantjetted from the refrigerant jet port 16 c of the nozzle 16. As a result,the configurations of respective parts in the ejector 16 can be easilyset, and thereby the ejector 16 can be easily formed.

The Other Embodiments

Although the present invention has been fully described in connectionwith the preferred embodiments thereof with reference to theaccompanying drawings, it is to be noted that various changes andmodifications will become apparent to those skilled in the art.

According to any one embodiment of the present invention, the ejector 16is provided with the suction passage 16 i through which the refrigerant(fluid) drawn from the refrigerant suction port 16 d flows into theinlet of the mixing and pressurizing portion 16 e. The passage sectionalarea of the suction passage 16 i is configured to be changed such thatthe refrigerant (fluid) drawn from the refrigerant suction port 16 d isdecompressed in the suction passage 16 i substantially in iso-entropy.Alternatively, the passage area of the suction passage 16 i isconfigured to be changed such that the flow velocity of the refrigerantflowing into the mixing and pressurizing portion 16 e from the suctionpassage 16 i is substantially equal to the flow velocity of therefrigerant (fluid) flowing from the jet port 16 c of the nozzle 16 ainto the mixing and pressurizing portion 16 e. Alternatively, thepassage sectional area of the suction passage 16 i is configured to bechanged such that the flow velocity of the fluid flowing into the mixingand pressurizing portion 16 e from the suction passage 16 i is equal toor larger than the sound velocity. In this case, the ejector efficiencycan be effectively improved. The other configurations in the ejector 16may be suitably changed or combined without being limited to theabove-described embodiments.

According to any one embodiment of the present invention, the mixing andpressurizing portion 16 e is configured by the straight portion 16 gextending from the inlet of the mixing and pressurizing portion 16 e ina range in the axial direction, and the expanding portion 16 h extendingcontinuously from a downstream end of the straight portion 16 g to theoutlet of the mixing and pressurizing portion 16 e. The straight portion16 g is a cylindrical passage having a constant passage area in itsentire range, and the expending portion 16 h is configured such that apassage sectional area of the expanding portion 16 h is graduallyincreased toward downstream in the flow direction of the refrigerant. Inthe ejector 16, the other configurations may be suitably changed orcombined without being limited to the above-described embodiments. Forexample, the range of the straight portion 16 g in the axial directionof the nozzle 16 a is set such that the flow velocities of gasrefrigerant and liquid refrigerant within the refrigerant flowing intothe mixing and pressurizing portion 16 e become equal to each other inthe range. Alternatively, when the length of the straight portion 16 gin the axial direction of the nozzle is L1 and the length from the inletof the mixing and pressurizing portion 16 e to the outlet of the mixingand pressurizing portion 16 e in the axial direction is L2, the mixingand pressurizing portion 16 e is configured such that 0<L1/L2≦0.4.Alternatively, the mixing and pressurizing portion 16 e may beconfigured such that the refrigerant is pressurized in iso-entropytherein.

In the above-described embodiments of the present invention, the ejector16 is used for the refrigeration cycle device 10 in which therefrigerant is branched at the branch portion 13 on an upstream side ofthe nozzle 16 a in the refrigerant flow from the radiator 12. However,the ejector 16 of the present invention can be used for a refrigerationcycle device without being limited to the examples of theabove-described embodiments.

For example, the ejector 16 of the present invention can be used for arefrigeration cycle device shown in FIG. 26. The refrigeration cycledevice shown in FIG. 26, an accumulator 20 is located downstream of theoutlet of the ejector 16 so that the refrigerant flowing out of theejector 16 can directly flow into the accumulator 20. The accumulator 20has a gas refrigerant outlet coupled to the refrigerant suction side ofthe compressor 11, and a liquid refrigerant outlet connected to arefrigerant inlet of an evaporator 19 so that the liquid refrigerantseparated from the gas refrigerant in the accumulator 20 flows into theevaporator 19. The gas refrigerant evaporated in the evaporator 19 isdrawn into the refrigerant suction port 16 d of the ejector 16. In therefrigeration cycle device shown in FIG. 26, the refrigerant flowingfrom the radiator 12 is decompressed in the nozzle 16 a and the gasrefrigerant from the evaporator 19 is drawn into the ejector 16 from therefrigerant suction port 16 d by the high-speed jet flow from the nozzle16 a. Even when the ejector 16 according to any one of the first totenth embodiments is used for the refrigeration cycle device shown inFIG. 26, the ejector efficiency can be improved.

In the example shown in FIGS. 5A and 5B, the suction gas refrigerant isdecompressed in iso-entropy in the suction passage 16 i; however, thesuction gas refrigerant is not limited to be decompressed iniso-entropy.

FIGS. 27A and 27B are modified examples of FIG. 5B. X and Y in FIG. 27Aand FIG. 27B correspond to the enlarged part VB in FIG. 5A. As shown inFIG. 27A, gas-liquid two-phase refrigerant can be drawn from therefrigerant suction port 16 b and can be decompressed in iso-entropy inthe suction passage 16 i of the ejector 16. Alternatively, as shown inFIG. 27B, the gas refrigerant is drawn from the refrigerant suction port16 d, and can be decompressed in iso-entropy into a gas-liquid two-phasesate.

In the above-described embodiments, the Freon-based refrigerant or thecarbon dioxide is typically used as the refrigerant. However, as therefrigerant, a generally-known refrigerant or a generally known fluidmay be used. For example, carbon-hydride based refrigerant may be usedas the refrigerant.

In the above-described embodiments, the refrigerant cycle device is usedfor a vehicle air conditioner or for a vehicle refrigerator. However,the refrigeration cycle device may be used for a fixed cooler, a fixedrefrigerator, a box having a cooling function, a cooling device for acoin machine or the like.

In the above-described embodiments, the first and second evaporators 17,19 are used as an interior heat exchanger for cooling air, and theradiator 12 is used as an exterior heat exchanger for radiating heat tooutside air. However, the first and second evaporators 17, 19 may beused as an exterior heat exchanger for absorbing heat from outside air,and the radiator 12 may be used as an interior heat exchanger forheating a fluid to be heated such as water or air. That is, the ejector16 of the present invention can be used for a heat pump cycle systemwith a heating function or/and a cooling function.

Such changes and modifications are to be understood as being within thescope of the present invention as defined by the appended claims.

1. An ejector comprising: a nozzle configured to decompress and expand afluid in any one state of a gas-liquid two-phase state, a liquid stateand a super-critical state; a body portion in which the nozzle isdisposed, the body portion having a fluid suction port from which afluid is drawn by a jet flow of the fluid jetted from a jet port of thenozzle, and a mixing and pressurizing portion in which the fluid jettedfrom the jet port of the nozzle and the fluid drawn from the fluidsuction port are mixed and kinetic energy of the mixed fluid in agas-liquid two-phase state is converted to pressure energy thereof; anda suction passage through which the fluid drawn from the fluid suctionport flows into an inlet of the mixing and pressurizing portion, whereina fluid passage area of the suction passage is configured to be changedsuch that the fluid drawn from the fluid suction port is decompressed inthe suction passage substantially in iso-entropy.
 2. An ejectorcomprising: a nozzle configured to decompress and expand a fluid in anyone state of a gas-liquid two-phase state, a liquid state and asuper-critical state; a body portion in which the nozzle is disposed,the body portion having a fluid suction port from which a fluid is drawnby a jet flow of the fluid jetted from a jet port of the nozzle, and amixing and pressurizing portion in which the fluid jetted from the jetport of the nozzle and the fluid drawn from the fluid suction port aremixed and kinetic energy of the mixed fluid in a gas-liquid two-phasestate is converted to pressure energy thereof; and a suction passagethrough which the fluid drawn from the fluid suction port flows into aninlet of the mixing and pressurizing portion, wherein a fluid passagearea of the suction passage is configured to be changed such that a flowvelocity of the fluid flowing into the mixing and pressurizing portionfrom the suction passage is substantially equal to a flow velocity ofthe fluid flowing from the jet port of the nozzle into the mixing andpressurizing portion.
 3. An ejector comprising: a nozzle configured todecompress and expand a fluid in any one state of a gas-liquid two-phasestate, a liquid state and a super-critical state; a body portion inwhich the nozzle is disposed, the body portion having a fluid suctionport from which a fluid is drawn by a jet flow of the fluid jetted froma jet port of the nozzle, and a mixing and pressurizing portion in whichthe fluid jetted from the jet port of the nozzle and the fluid drawnfrom the fluid suction port are mixed and kinetic energy of the mixedfluid in a gas-liquid two-phase state is converted to pressure energythereof; and a suction passage through which the fluid drawn from thefluid suction port flows into an inlet of the mixing and pressurizingportion, wherein a fluid passage area of the suction passage isconfigured to be changed such that a flow velocity of the fluid flowinginto the mixing and pressurizing portion from the suction passage isequal to or larger than a sound velocity.
 4. The ejector according toclaim 1, wherein the fluid passage area of the suction passage isgradually reduced toward downstream in a flow direction of the fluidflowing in the suction passage.
 5. The ejector according to claim 4,wherein a reduce degree of the fluid passage area at an inlet side ofthe suction passage is larger than a reduce degree of the fluid passagearea at an outlet side of the suction passage.
 6. The ejector accordingto claim 1, wherein the fluid passage area of the suction passage at aninlet side of the suction passage is gradually reduced toward downstreamin a flow direction of the fluid flowing in the suction passage, and thefluid passage area of the suction passage at an outlet side of thesuction passage is gradually increased toward downstream in the flowdirection of the fluid flowing in the suction passage.
 7. The ejectoraccording to claim 1, wherein the suction passage is provided between anouter peripheral surface of the nozzle and an inner peripheral surfaceof the body portion.
 8. The ejector according to claim 1, wherein thesuction passage is configured by another nozzle to be provided therein.9. The ejector according to claim 1, wherein the nozzle and the suctionpassage are configured, such that an enthalpy difference (ΔH) betweenenthalpy of the fluid at an inlet of the nozzle and enthalpy of thefluid at the jet port of the nozzle is equal to or larger than anenthalpy difference (Δh) between enthalpy of the fluid at the inlet ofthe suction passage and enthalpy of the fluid at the outlet of thesuction passage.
 10. The ejector according to claim 1, wherein themixing and pressurizing portion is configured by a straight portionextending from the inlet of the mixing and pressurizing portion in arange, and an expanding portion extending from a downstream end of thestraight portion to the outlet of the mixing and pressurizing portion,the straight portion is cylindrical passage having a constant passagearea in its entire range, and the expending portion is configured suchthat a passage sectional area of the expanding portion is graduallyincreased toward downstream in a flow direction of the fluid.
 11. Theejector according to claim 10, wherein the range of the straight portionis set such that the flow velocities of gas fluid and liquid fluidwithin the fluid flowing into the mixing and pressurizing portion becomeequal to each other in the range.
 12. The ejector according to claim 10,wherein when a length of the straight portion in an axial direction ofthe nozzle is L1 and a length from the inlet of the mixing andpressurizing portion to the outlet of the mixing and pressurizingportion in the axial direction is L2, the mixing and pressurizingportion is configured such that 0<L1/L2≦0.4.
 13. The ejector accordingto claim 10, wherein the mixing and pressurizing portion is configuredsuch that the fluid is pressurized in iso-entropy.
 14. An ejectorcomprising: a nozzle configured to decompress and expand a fluid in anyone state of a gas-liquid two-phase state, a liquid state and asuper-critical state; and a body portion in which the nozzle isdisposed, the body portion having a fluid suction port from which afluid is drawn by a jet flow of the fluid jetted from a jet port of thenozzle, and a mixing and pressurizing portion in which the fluid jettedfrom the jet port of the nozzle and the fluid drawn from the fluidsuction port are mixed and kinetic energy of the mixed fluid in agas-liquid two-phase state is converted to pressure energy thereof,wherein the mixing and pressurizing portion is configured by a straightportion extending from the inlet of the mixing and pressurizing portionin a range, and an expanding portion extending from a downstream end ofthe straight portion to the outlet of the mixing and pressurizingportion, the straight portion is a cylindrical passage having a constantpassage area in its entire range, and the expending portion isconfigured such that a passage sectional area of the expanding portionis gradually increased toward downstream in a flow direction of thefluid.
 15. The ejector according to claim 14, wherein the range of thestraight portion is set such that the flow velocities of gas fluid andliquid fluid within the fluid flowing into the mixing and pressurizingportion become equal to each other in the range.
 16. The ejectoraccording to claim 14, wherein when a length of the straight portion inan axial direction of the nozzle is L1 and a length from the inlet ofthe mixing and pressurizing portion to the outlet of the mixing andpressurizing portion in the axial direction is L2, the mixing andpressurizing portion is configured such that 0<L1/L2≦0.4.
 17. Theejector according to claim 14, wherein the mixing and pressurizingportion is configured such that the fluid is pressurized in iso-entropyin the mixing and pressurizing portion.
 18. The ejector according toclaim 14, wherein a sectional shape of a wall surface of the expandingportion in a section including an axial line of the nozzle is a straightline.
 19. The ejector according to claim 14, wherein a sectional shapeof a wall surface of the expanding portion in a section including anaxial line of the nozzle is a curved line.
 20. The ejector according toclaim 14, wherein a sectional shape of a wall surface of the expandingportion in a section including an axial line of the nozzle is formed bycombining plural straight lines.
 21. The ejector according to claim 14,wherein a sectional shape of a wall surface of the expanding portion ina section including an axial line of the nozzle is formed by combiningat least a straight line and a curved line.
 22. The ejector according toclaim 14, wherein an expanding degree of the expanding portion at aninlet side of the expanding portion is larger than an expanding degreeof the expanding portion at an outlet side of the expanding portion. 23.The ejector according to claim 2, wherein the fluid passage area of thesuction passage is gradually reduced toward downstream in a flowdirection of the fluid flowing in the suction passage.
 24. The ejectoraccording to claim 23, wherein a reduce degree of the fluid passage areaat an inlet side of the suction passage is larger than a reduce degreeof the fluid passage area at an outlet side of the suction passage. 25.The ejector according to claim 2, wherein the fluid passage area of thesuction passage at an inlet side of the suction passage is graduallyreduced toward downstream in a flow direction of the fluid flowing inthe suction passage, and the fluid passage area of the suction passageat an outlet side of the suction passage is gradually increased towarddownstream in the flow direction of the fluid flowing in the suctionpassage.
 26. The ejector according to claim 2, wherein the suctionpassage is provided between an outer peripheral surface of the nozzleand an inner peripheral surface of the body portion.
 27. The ejectoraccording to claim 2, wherein the suction passage is configured byanother nozzle to be provided therein.
 28. The ejector according toclaim 2, wherein the nozzle and the suction passage are configured, suchthat an enthalpy difference (ΔH) between enthalpy of the fluid at aninlet of the nozzle and enthalpy of the fluid at the jet port of thenozzle is equal to or larger than an enthalpy difference (Δh) betweenenthalpy of the fluid at the inlet of the suction passage and enthalpyof the fluid at the outlet of the suction passage.
 29. The ejectoraccording to claim 3, wherein the fluid passage area of the suctionpassage is gradually reduced toward downstream in a flow direction ofthe fluid flowing in the suction passage.
 30. The ejector according toclaim 29, wherein a reduce degree of the fluid passage area at an inletside of the suction passage is larger than a reduce degree of the fluidpassage area at an outlet side of the suction passage.
 31. The ejectoraccording to claim 3, wherein the fluid passage area of the suctionpassage at an inlet side of the suction passage is gradually reducedtoward downstream in a flow direction of the fluid flowing in thesuction passage, and the fluid passage area of the suction passage at anoutlet side of the suction passage is gradually increased towarddownstream in the flow direction of the fluid flowing in the suctionpassage.
 32. The ejector according to claim 3, wherein the suctionpassage is provided between an outer peripheral surface of the nozzleand an inner peripheral surface of the body portion.
 33. The ejectoraccording to claim 3, wherein the suction passage is configured byanother nozzle to be provided therein.
 34. The ejector according toclaim 3, wherein the nozzle and the suction passage are configured, suchthat an enthalpy difference (ΔH) between enthalpy of the fluid at aninlet of the nozzle and enthalpy of the fluid at the jet port of thenozzle is equal to or larger than an enthalpy difference (Δh) betweenenthalpy of the fluid at the inlet of the suction passage and enthalpyof the fluid at the outlet of the suction passage.